U.S. patent number 7,971,449 [Application Number 11/660,170] was granted by the patent office on 2011-07-05 for heat-activated heat-pump systems including integrated expander/compressor and regenerator.
This patent grant is currently assigned to N/A, State of Oregon acting by and through the State Board of Higher Education on Behalf of Oregon State University. Invention is credited to Monte K. Drost, Thomas G. Herron, Richard B. Peterson.
United States Patent |
7,971,449 |
Peterson , et al. |
July 5, 2011 |
Heat-activated heat-pump systems including integrated
expander/compressor and regenerator
Abstract
Heat-activated heat-pump systems and related methods are
disclosed that include a power cycle coupled to a vapor-compression
refrigeration cycle both utilizing the same working fluid. The
power cycle comprises a boiler, an expander receiving superheated
vapor and producing work from the superheated vapor, a condenser,
and a pump. A regenerator conducts a first stream of working fluid
from the pump to the boiler and a second stream of the working
fluid from the expander to the condenser while transferring heat
from the second stream to the first stream. The refrigeration cycle
comprises a compressor that compresses the working fluid from the
evaporator and delivers the compressed working fluid to a
condenser. The expander and compressor are coupled together such
that at least a portion of the work produced by the expander is
utilized for running the compressor.
Inventors: |
Peterson; Richard B.
(Corvallis, OR), Herron; Thomas G. (Bend, OR), Drost;
Monte K. (Corvallis, OR) |
Assignee: |
State of Oregon acting by and
through the State Board of Higher Education on Behalf of Oregon
State University (Corvallis, OR)
N/A (N/A)
|
Family
ID: |
37637629 |
Appl.
No.: |
11/660,170 |
Filed: |
August 15, 2005 |
PCT
Filed: |
August 15, 2005 |
PCT No.: |
PCT/US2005/029112 |
371(c)(1),(2),(4) Date: |
February 13, 2007 |
PCT
Pub. No.: |
WO2007/008225 |
PCT
Pub. Date: |
January 18, 2007 |
Prior Publication Data
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Document
Identifier |
Publication Date |
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US 20080006040 A1 |
Jan 10, 2008 |
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Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
Issue Date |
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60601478 |
Aug 14, 2004 |
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Current U.S.
Class: |
62/324.1;
62/116 |
Current CPC
Class: |
F04B
35/00 (20130101); B60H 1/3204 (20130101); F01K
21/00 (20130101); F01K 13/00 (20130101); G01M
3/26 (20130101); F25B 27/02 (20130101); F01K
25/08 (20130101); F25B 2400/14 (20130101); Y02A
30/274 (20180101) |
Current International
Class: |
F25B
13/00 (20060101) |
Field of
Search: |
;62/116,323.1,324.1,324.6 |
References Cited
[Referenced By]
U.S. Patent Documents
Other References
Garris, Charles A., Jr. et al., "The Pressure-Exchange Ejector Heat
Pump," AES-vol. 38, Proceedings of the ASME Advanced Energy Systems
Division, ASME 1998. cited by other .
Huang, B.J. et al., "Ejector Performance Characteristics and Design
Analysis of Jet Refrigeration System," vol. 107, Transactions of
the ASME, Jul. 1985, pp. 792-802. cited by other .
Cahill, J.W., "A Novel Engine and Heat Pump," www.eigenmorph.com,
undated, pp. 1-5. cited by other.
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Primary Examiner: Jones; Melvin
Attorney, Agent or Firm: Klarquist Sparkman, LLP
Government Interests
ACKNOWLEDGMENT OF GOVERNMENT SUPPORT
This invention was developed under contract no. CFDA 12.910 from
the U.S. Department of Defense. The U.S. government has certain
rights in this invention.
Parent Case Text
CROSS REFERENCE TO RELATED APPLICATIONS
This is the U.S. National Stage of International Application No.
PCT/US2005/029112, filed Aug. 15, 2005, which was published in
English under PCT Article 21(2), which in turn claims the benefit
of U.S. Provisional Application No. 60/601,478, filed Aug. 14,
2004. Both applications are incorporated herein in their entirety.
Claims
What is claimed is:
1. A heat-activated heat-pump system, comprising a power cycle
coupled to a vapor-compression refrigeration cycle both utilizing
the same working fluid; the power cycle comprising a boiler that
vaporizes and superheats the working fluid, an expander coupled to
receive superheated vapor from the boiler and configured to produce
work from the superheated vapor, a condenser coupled to receive
working fluid from the expander and configured to condense the
working fluid, a pump coupled to receive condensed working fluid
from the condenser and to return the working fluid to the boiler,
and a regenerator coupled to conduct a first stream of the working
fluid from the pump to the boiler and to conduct a second stream of
the working fluid from the expander to the condenser while
transferring heat from the second stream to the first stream; the
refrigeration cycle comprising the condenser, a pressure-reducing
throttling valve, an evaporator coupled to receive reduced-pressure
working fluid via the throttling valve from the condenser and
configured to expand the working fluid sufficiently to cool the
working fluid, and a compressor configured to compress the working
fluid from the evaporator and to deliver the compressed working
fluid to the condenser; and the expander and compressor being
coupled together such that at least a portion of the work produced
by the expander is utilized for running the compressor, the
expander being configured to achieve expansion of the superheated
vapor of the working fluid substantially adiabatically and
substantially isentropically.
2. The system of claim 1, wherein the power cycle is a Rankine
power cycle.
3. The system of claim 1, wherein the pump is coupled to the
expander such that a portion of the work produced by the expander
is utilized for running the pump.
4. The system of claim 1, wherein the expander and the compressor
are integrated with each other.
5. The system of claim 4, wherein: the expander comprises a shaft
that is rotated by work produced in the expander; the compressor
comprises a shaft; and the shaft of the expander is coupled to the
shaft of the compressor.
6. The system of claim 4, wherein the integrated expander and
compressor exhibit an isentropic efficiency of at least 70%.
7. The system of claim 1, further comprising a combustor coupled to
the boiler, the combustor being configured to combust a fuel to
generate heat sufficient to vaporize and superheat the working
fluid, and to supply the heat to the working fluid in the
boiler.
8. The system of claim 7, wherein the fuel is a hydrocarbon
fuel.
9. The system of claim 7, wherein the combustor and the boiler are
integrated with each other such that heat produced in the combustor
is transferred in the boiler directly to the working fluid.
10. The system of claim 1, wherein the working fluid is
isopentane.
11. The system of claim 1, wherein the working fluid is a
fluorinated-hydrocarbon refrigerant.
12. The system of claim 1, further comprising a coolant cycle
coupled to the evaporator, the coolant cycle being configured to
urge flow of a coolant from the evaporator, in which the coolant
surrenders heat to the working fluid, to a cooling zone at which
the coolant absorbs heat from a region in thermal contact with the
cooling zone.
13. The system of claim 12, wherein: the coolant is water; and the
cooling zone is a water chest in thermal contact with a person's
body so as to cool the person's body.
14. The system of claim 13, wherein the cooling zone is a
heat-exchanger configured to cool vehicular air in thermal contact
with the heat-exchanger.
15. The system of claim 1, further comprising at least one
balance-of-plant component coupled to the expander in a manner
allowing the balance-of-plant component to capture and utilize a
portion of the work produced by the expander.
16. The system of claim 1, wherein at least one of the boiler, the
evaporator, the condenser, and the regenerator comprises
microfluidic channels for conducting the working fluid and for
conducting heat relative to the working fluid.
17. The system of claim 1, wherein at least one of the boiler, the
evaporator, the condenser, and the regenerator comprises
microtubules for conducting the working fluid and for conducting
heat relative to the working fluid.
18. The system of claim 1, wherein at least one of the expander and
compressor is positive-displacement.
19. The system of claim 1, wherein both the expander and compressor
are positive-displacement.
20. The system of claim 19, wherein each of the expander and
compressor comprises at least one respective piston that is movable
relative to a respective cylinder.
21. The system of claim 20, wherein: the expander comprises two
respective pistons in respective cylinders; and the compressor
comprises two respective pistons in respective cylinders.
22. The system of claim 21, wherein: the expander pistons operate
180.degree. out of phase with each other; and the compressor
pistons operate 180.degree. out of phase with each other.
23. The system of claim 22, wherein the expander pistons and the
compressor pistons are coupled to a common shaft that is rotated by
translation of reciprocating motion of the expander pistons to the
shaft.
24. The system of claim 21, wherein: the expander comprises, for
each cylinder thereof, a respective inlet valve and respective
outlet valve that control flow of working fluid into and out of the
cylinder; and the inlet valves and outlet valves are actuated in a
self-timed manner in coordination with reciprocating motion of the
respective actuator pistons.
25. The system of claim 24, wherein the inlet valves and outlet
valves of the expander comprise respective poppets that are
actuated by reciprocating motion of the respective pistons in the
cylinders.
26. The system of claim 25, wherein: each expander piston comprises
a respective tappet for the respective inlet valve and outlet
valve; and during reciprocating motion of the expander pistons, the
tappets engage respective poppets of the respective inlet and
outlet valves.
27. The system of claim 21, wherein: the compressor comprises, for
each cylinder thereof, a respective inlet valve and respective
outlet valve that control flow of working fluid into and out of the
cylinder; and the inlet valves and outlet valves are actuated in a
self-timed manner in coordination with reciprocating motion of the
respective compressor pistons.
28. The system of claim 27, wherein the inlet valves and outlet
valves of the compressor comprise respective flappers that open and
close automatically at respective times during reciprocating motion
of the respective compressor pistons.
29. The system of claim 1, wherein the boiler is configured to heat
the working fluid at substantially constant pressure.
30. A vehicle comprising an air-conditioning system including a
heat-pump system as recited in claim 1 situated and configured to
adjust the temperature of air in the vehicle.
31. A suit configured to be worn by a person, the suit comprising a
heat-pump system as recited in claim 1 situated and configured to
adjust the temperature of the person wearing the suit.
32. In a heat-activated heat-pump system including a power cycle
coupled to a vapor-compression refrigeration cycle both utilizing
the same working fluid, an integrated expander/compressor,
comprising: an expander situated in the power cycle and that
produces work from superheated vapor of the working fluid
introduced to the expander, the expander comprising a first pair of
movable pistons disposed in respective stationary cylinders, the
pistons being coupled to a shaft and configured to move
reciprocatingly, while rotating the shaft, 180.degree. out of phase
with each other, each of the cylinders comprising respective inlet
and outlet valves that are actuated by movement of the respective
piston; and a compressor situated in the refrigeration cycle and
that compresses the working fluid from the evaporator, the
compressor comprising a second pair of movable pistons disposed in
respective stationary cylinders, the pistons being coupled to the
shaft such that at least a portion of the work produced by the
expander is utilized for running the compressor, the pistons of the
compressor and configured to move reciprocatingly, with rotation of
the shaft, 180.degree. out of phase with each other; each of the
cylinders comprising respective inlet and outlet valves that are
actuated by movement of the respective piston.
33. The system of claim 32, wherein the inlet valves and outlet
valves of the expander comprise respective poppets that are
actuated by reciprocating motion of the respective pistons in the
cylinders.
34. The system of claim 32, wherein: each expander piston comprises
a respective tappet for the respective inlet valve and outlet
valve; and during reciprocating motion of the expander pistons, the
tappets engage respective poppets of the respective inlet and
outlet valves.
35. The system of claim 32, wherein the inlet valves and outlet
valves of the compressor are actuated in a self-timed manner in
coordination with reciprocating motion of the respective compressor
pistons.
36. The system of claim 35, wherein the inlet valves and outlet
valves comprise respective flappers that open and close
automatically at respective times during reciprocating motion of
the respective compressor pistons.
37. A cooling system, comprising: a refrigeration cycle coupled to
a power cycle that drives the refrigeration cycle and uses a same
working fluid as the refrigeration cycle; the power cycle
comprising a boiler that receives waste heat by which the boiler
vaporizes and superheats the working fluid, an expander that
receives superheated vapor from the boiler and produces work from
the superheated vapor, a condenser that receives working fluid from
the expander and condenses the working fluid, a pump that receives
condensed working fluid from the condenser and returns the working
fluid to the boiler, and a regenerator that conducts a first stream
of the working fluid from the pump to the boiler and a second
stream of the working fluid from the expander to the condenser
while transferring heat from the second stream to the first stream;
the refrigeration cycle comprising the condenser, an evaporator
that receives working fluid from the condenser and internally
evaporates the working fluid in a manner that cools the working
fluid in the evaporator, a compressor that compresses the working
fluid from the evaporator and delivers the compressed working fluid
to the condenser, and a throttling valve that delivers working
fluid from the condenser to the evaporator while reducing the
pressure of the working fluid; and the expander and compressor
being coupled together such that at least a portion of the work
produced by the expander is utilized directly for running the
compressor, and the expander is configured to achieve expansion of
the superheated vapor of the working fluid substantially
adiabatically and substantially isentropically.
38. The cooling system of claim 37, further comprising a
circulation loop to and from the evaporator, the circulation loop
conducting a liquid that is circulated to the evaporator for
cooling of the liquid.
39. The cooling system of claim 38, wherein the cooled liquid is
circulated to a cooling zone in thermal contact with a thing so to
cool the thing.
40. The cooling system of claim 39, wherein the cooling zone is in
contact with a person so as to cool the person.
41. The cooling system of claim 39, wherein the cooling zone is
contacted by air in a vehicle so as to cool the air in the
vehicle.
42. A heat-activated heat-pump system, comprising power-cycle means
coupled to refrigeration-cycle means both utilizing the same
working fluid; said power-cycle means comprising superheating means
for vaporizing and superheating the working fluid, work-producing
means for producing work from superheated vapor received from said
superheating means, condensing means for condensing the working
fluid received from said work-producing means, pump means for
delivering condensed working fluid from said condensing means to
said superheating means, and regenerator means for transferring
heat, from a first stream of working fluid flowing from said
work-producing means to said condensing means, to a second stream
of working fluid flowing from said pump means to said superheating
means; said refrigeration-cycle means comprising said condensing
means, pressure-reducing means for reducing pressure of the working
fluid from said condensing means, evaporator means for expanding
working fluid, received via said pressure-reducing means from said
condensing means, sufficiently to cool the working fluid, and
compressing means for compressing working fluid received from said
evaporator means for delivery to said condensing means; and said
work-producing means and said compressing means being coupled
together such that at least a portion of the work produced by said
work-producing means is utilized for running said compressing
means.
43. The system of claim 42, further comprising combusting means for
combusting a fuel to produce waste heat and for supplying waste
heat to said superheating means sufficiently for the superheating
means to vaporize and superheat the working fluid.
44. The system of claim 43, wherein said combusting means and said
superheating means are integrated with each other.
45. The system of claim 42, wherein said work-producing means
comprises expansion means for expanding the superheated working
fluid in a manner by which work is produced.
46. The system of claim 42, wherein said work-producing means and
said compressing means are integrated with each other.
47. The system of claim 42, further comprising coolant-cycle means
for circulating a coolant from said evaporating means to a cooling
zone.
48. The system of claim 47, wherein the cooling zone is in thermal
contact with a body of a person.
49. The system of claim 47, wherein the cooling zone is in thermal
contact with air in a vehicle interior.
50. The system of claim 42, wherein said work-producing means and
said compressing means each comprise positive-displacement means
for moving the working fluid through said work-producing means and
compressing means.
51. A protective-suit means, comprising a system as recited in
claim 42, the system being coupled to said protective-suit means in
a manner by which the system cools said protective-suit means.
52. A method for removing heat from a body, comprising: thermally
contacting the body with a coolant such that the coolant picks up
and removes heat from the body; circulating the coolant through an
evaporator, of a refrigeration cycle utilizing a working fluid, so
as to contact the coolant thermally with working fluid cooled in
the evaporator; circulating a first portion of the working fluid
through the refrigeration cycle so as to remove heat from the
working fluid, the refrigeration cycle including a compressor
situated downstream of the evaporator; and circulating a second
portion of the working fluid through a power cycle including a
boiler, an expander situated downstream of the boiler, a first
compartment of a regenerator situated upstream of the boiler, and a
second compartment of the regenerator situated downstream of the
expander, such that heat exchange occurs in the regenerator between
the first and second compartments.
53. The method of claim 52, wherein the step of circulating the
first portion of the working fluid through the refrigeration cycle
further comprises passing the first portion of the working fluid
through a compressor situated downstream of the evaporator.
54. The method of claim 53, wherein circulating the second portion
of the working fluid through the expander produces work, the method
further comprising coupling the compressor to the expander such
that at least a portion of the work is utilized for operating the
compressor.
55. A method for removing heat from a body, comprising: thermally
contacting the body with a coolant such that the coolant picks up
heat from the body; circulating the coolant through an evaporator,
of a refrigeration cycle utilizing a working fluid, so as to
contact the coolant thermally with working fluid cooled in the
evaporator; circulating a first portion of the working fluid
through the evaporator to cool the working fluid and thus the
coolant, through a compressor to compress the working fluid,
through a condenser to condense and heat the working fluid, and
back to the evaporator; circulating a second portion of the working
fluid from the condenser through a first compartment of a
regenerator to a boiler, through the boiler to superheat the
working fluid, through an expander to expand the superheated
working fluid and to extract and utilize work from the superheated
working fluid, and through a second compartment of the regenerator
back to the condenser; combining the second portion of the working
fluid from the second compartment of the regenerator with the first
portion to produce a combined stream of working fluid entering the
condenser; and in the regenerator, transferring heat from the
working fluid in the second compartment to the working fluid in the
first compartment.
56. The method of claim 55, further comprising: combusting a fuel
in a combustor to produce heat; and providing the heat to the
boiler.
57. The method of claim 55, further comprising coupling the
expander to the compressor so as to utilize at least a portion of
the work extracted by the expander for running the compressor.
58. The method of claim 55, wherein the body is of a person.
59. The method of claim 55, wherein the body is of air.
Description
FIELD
This disclosure is directed to, inter alia, heat-activated
thermodynamic cycles and heat-pump systems that include a power
cycle and a refrigeration cycle. The systems include an integrated
expander/compressor and at least one "regenerator"
(heat-exchanger), and are especially suitable for use in, for
example, compact and light-weight cooling units for vehicles and
individual personnel.
BACKGROUND
Combustion-driven, heat-activated heat pumps used for heating
and/or cooling have a large performance advantage in terms of size
and weight over battery-powered heat-pumping devices. This is due
in part to the respective energy densities of commonly used
liquid-hydrocarbon fuels (in the vicinity of 42 kJ/g for JP-8 and
diesel fuel) compared to the energy densities of zinc/air batteries
(approximately 1.2 kJ/g) and of lead-acid batteries (approximately
0.12 kJ/g). High-performance, heat-activated cooling systems able
to exploit this advantage of hydrocarbon fuels (by combusting them)
would have many commercial and military applications such as
cooling of personnel-protective suits (e.g., chemical- and/or
biological-protective suits), cooling of vehicle interiors, and
recovering and using waste heat from other processes. Even with a
heat-to-work conversion efficiency of 10 to 20%, a
combustion-driven heat-activated cooling system would be smaller
and lighter, and could operate for longer periods of time (compared
to battery-powered units) if component size and weight could be
effectively limited.
Heat-activated heat-pumps are similar to conventional
vapor-compression heat-pumps in that both utilize a working fluid
and both include a compressor. In general, the primary difference
between a heat-activated heat-pump and a vapor-compression
heat-pump is the manner in which compression of the working fluid
is accomplished, or in the manner in which power is supplied to the
compressor. For example, a classic heat-activated refrigeration
process is utilizes a jet-ejector cycle. Although a jet-ejector
cycle is simple in design, generally reliable, and able to utilize
waste heat, this cycle has not found wide-spread application
because it exhibits poor thermodynamic performance. Also, the
efficiency of these systems is poor. For example, the heat-activate
coefficient of performance (COP), defined as the amount of cooling
provided by the cycle divided by the amount of heat required to
drive the cycle, is usually very low, typically less than 0.3.
Also, the efficiency of these devices diminishes with decreasing
system size. For a portable system, a low COP not only increases
the size and weight of the boiler and condenser in the Rankine
portion of the cycle, but it also increases the weight and volume
of fuel that must be carried.
BACKGROUND REFERENCES INCLUDE
Drost et al., 1998, "Miniature Heat Pumps for Portable and
Distributed Space Conditioning Applications," AIChE 1998 Spring
National Meeting, New Orleans; Drost et al., 1999, "Mesoscopic
Heat-Actuated Heat Pump Development," ASME IMECE Conference,
Nashville, Tenn.; Kouremenos et al., 1998, "Optimization of
Enhanced Steam-Ejector Applied to Steam Jet Refrigeration,"
Proceedings of the ASME Advanced Energy Systems Division, AES-Vol.
38; Huang et al., 1985, "Ejector Performance Characteristics and
Design Analysis of Jet Refrigeration System," J. Eng. Gas. Turbines
and Power, ASME Transactions 107(3):792-802; Lee et al., "Influence
of Cyclic Wall-to-Gas Heat Transfer in the Cylinder of the Valved
Hot-Gas Engine," Proceedings of the 13th Intersociety Energy
Conversion Engineering Conference, 1978, pp. 1798-1804; Granet et
al., Thermodynamics and Heat Power, 6th ed., Prentice-Hall, New
Jersey, 2000, pp. 275-278.
SUMMARY
The deficiencies of conventional systems are addressed by systems
and methods as disclosed herein.
According to a first aspect, heat-activated heat-pump systems are
provided that comprise a power cycle coupled to a vapor-compression
refrigeration cycle, wherein both cycles utilize the same working
fluid. In an embodiment of such a system the power cycle comprises
a boiler that vaporizes and superheats the working fluid, an
expander coupled to receive superheated vapor from the boiler and
configured to produce work from the superheated vapor, a condenser
coupled to receive working fluid from the expander and configured
to condense the working fluid, and a pump coupled to receive
condensed working fluid from the condenser and to return the
working fluid to the boiler. The power cycle also includes a
regenerator coupled to conduct a first stream of the working fluid
from the pump to the boiler and to conduct a second stream of the
working fluid from the expander to the condenser while transferring
heat from the second stream to the first stream. The refrigeration
cycle comprises the condenser, a pressure-reducing throttling
valve, an evaporator coupled to receive reduced-pressure working
fluid via the throttling valve from the condenser and configured to
expand the working fluid sufficiently to cool the working fluid,
and a compressor configured to compress the working fluid from the
evaporator and to deliver the compressed working fluid to the
condenser. The expander and compressor are coupled together such
that at least a portion of the work produced by the expander is
utilized for running the compressor. The refrigeration cycle is
essentially a vapor-compression cycle, and an exemplary power cycle
in this system is a Rankine power cycle. By expanding the working
fluid in the expander, the work generated by the expansion is
effectively utilized for driving the compressor. Since the work
output from the power cycle is used to drive the refrigeration
cycle, an overall cycle performance is achieved that readily can
exceed the performance of a conventional jet-ejector cycle.
The working fluid can be any suitable fluid capable of assuming
liquid and vapor states at appropriate locations and times in the
cycle under the conditions of use of the cycle. By way of example,
the working fluid can be any of various hydrocarbons such as
isopentane, any of various fluorocarbon refrigerants, or any of
various other suitable working fluids.
The expander and the compressor desirably are at least coupled to
each other. For example, the expander can comprise a shaft that is
rotated by work produced in the expander, and the compressor can
comprise a shaft. The shaft of the expander desirably is coupled to
the shaft of the compressor so that, as the expander shaft rotates,
corresponding rotation of the compressor shaft occurs. Further
desirably, the expander and compressor are not only coupled to each
other but also integrated with each other, such as in the same
housing, to minimize thermal loss between these two components. A
desirable performance standard is for the integrated expander and
compressor to exhibit an isentropic efficiency of at least 70%.
The system also desirably further comprises a combustor coupled to
the boiler, wherein the combustor is configured: (a) to combust a
fuel to generate heat sufficient to vaporize and superheat the
working fluid, and (b) to supply the heat to the working fluid in
the boiler. The fuel can be any suitable fuel such as a hydrocarbon
fuel. Further desirably, the combustor and the boiler are
integrated with each other such that heat produced in the combustor
is transferred in the boiler directly to the working fluid. By
exploiting the higher stored-energy density of, for example, liquid
hydrocarbons over batteries, this heat-pump cycle has a large
performance advantage (size, weight, and portability) over
conventional battery-powered vapor-compression systems.
The system further can comprise a coolant cycle coupled to the
evaporator. In an embodiment the coolant cycle is configured to
urge flow of a coolant from the evaporator, in which the coolant
surrenders heat to the working fluid, to a cooling zone at which
the coolant absorbs heat from a region in thermal contact with the
cooling zone. (As used herein, "thermal contact" includes both
direct contact and indirect contact. In either event, the contact
is sufficient to achieve conduction of heat between the bodies in
thermal contact with each other, so the conduction can be via an
intermediate body.) By way of example, the coolant is water. This
coolant can be used in an arrangement in which the cooling zone is
a water chest in thermal contact with a person's body; thus, the
circulating coolant cools the person's body. Such an application is
especially advantageous in certain types of protective suits worn
by a person, such as a chemical-protection suit or
biological-isolation suit. In another application the cooling zone
can be a heat-exchanger configured to cool vehicular air in thermal
contact with the heat-exchanger. Such an application is especially
advantageous for use in motor vehicles, especially as used in hot
climates or other elevated-temperature conditions.
One or more of the heat-exchange components of the system (e.g.,
combustor/boiler, regenerator, condenser, and evaporator) can be
configured using microtechnology-based structures to enhance heat
transfer and to reduce system weight and volume. Computational
models were developed and executed, as described herein, to
estimate system performance for a set of given components and
operating conditions. Results of one study revealed attainment of a
heat-activate coefficient of performance (COP) of 1.3 with
100.degree. C. of superheat over a saturation temperature of
116.degree. C. This is a very good COP for a heat-activated
system.
The foregoing and additional features and advantages of the subject
systems and methods will be more readily apparent from the
following detailed description, which proceeds with reference to
the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a schematic diagram of an exemplary embodiment of a
heat-pump system (also termed a "cycle") including heat
recovery.
FIG. 2 is a T-S diagram for the system of FIG. 1, according to the
conditions set forth in Table 1. The T-S diagram includes
respective portions for the power cycle and the refrigeration
cycle.
FIG. 3 is a plot, for the system of FIG. 1, of coefficient of
performance (COP) versus boiler temperature, revealing the effect
of the regenerator on the COP of the total system whenever the
pressure of the boiler remains at 1000 kPa and superheat is
increased from zero (at a boiler temperature of 116.degree. C.) to
227.degree. C.
FIG. 4 is a plot, for the system of FIG. 1, of COP versus expander
efficiency, showing that the COP decreases almost linearly with a
corresponding decrease in expander efficiency.
FIG. 5 is a plot, for the system of FIG. 1, showing the
relationship of system weight to condenser temperature.
FIG. 6 is a plot, for the system of FIG. 1, showing the
relationship of COP to condenser temperature.
FIG. 7 is a plot, for the system of FIG. 1, showing the influence
of evaporator temperature on the overall system COP.
FIGS. 8(a)-8(c) schematically depict, for the first representative
embodiment of the expander, three respective steps of a piston
cycle. FIG. 8(a) depicts the start of a downward stroke, when the
piston is at top-dead-center (TDC); FIG. 8(b) depicts the piston
that has traveled about 30% of the downward stroke; and FIG. 8(c)
depicts the piston at the end of the downward stroke, when the
piston is at bottom-dead-center (BDC).
FIG. 9(a) schematically depicts certain details of the two pistons
and their respective cylinders of the first representative
embodiment of the expander.
FIG. 9(b) is an elevational section depicting certain details of
the upper portion of a piston, cylinder, cylinder head, inlet
valve, and exhaust valve of the first representative embodiment of
the expander.
FIG. 10 is an elevational section of a portion of the cap of a
piston of the expander according to the first representative
embodiment. The cap defines a seal comprising a flexible lip that
engages the inside wall of the cylinder.
FIG. 11 is a schematic diagram showing a basic thermodynamic model
of the first representative embodiment of the expander.
FIG. 12 is a more detailed schematic diagram of the thermodynamic
model shown in FIG. 11.
FIG. 13 is a schematic diagram showing the process-flow measurement
setup used for performing evaluations of the first representative
embodiment of the expander.
FIG. 14 is a schematic diagram showing the measurement setup used
for producing pressure-volume curves of the expander cycle obtained
using the first representative embodiment of the expander.
FIG. 15 is a plot of shaft power as a function of shaft rotational
velocity for various inlet pressures to the first representative
embodiment of the expander.
FIG. 16 is a plot of torque data as a function of shaft rotational
velocity for various inlet pressures to the first representative
embodiment of the expander, showing that torque is relatively
independent of shaft rotational velocity.
FIG. 17 is a plot of calculated isentropic efficiency, of the first
representative embodiment of the expander, as a function of shaft
rotational speed for inlet pressures ranging from 35 psia to 75
psia (20 psig to 60 psig).
FIG. 18 is a plot of polytropic efficiency, of the first
representative embodiment of the expander, as a function of shaft
velocity for inlet pressures ranging from 35 psia to 75 psia. Heat
transfer was calculated using Equation (13).
FIG. 19 is a plot, for the first representative embodiment of the
expander, of P-V (pressure-volume) data recorded with the expander
operating at 1500 rpm shaft speed and at an inlet pressure ranging
from 35 psia to 75 psia.
FIG. 20(a) schematically depicts certain details of the pistons,
cylinders, and flapper valves of the first representative
embodiment of the compressor.
FIG. 20(b) is an exploded view showing more specific details of the
flapper valves used in the first representative embodiment of the
compressor.
FIG. 21 is an elevational section of the upper portion of one of
the cylinders of an expander according to a second representative
embodiment.
FIG. 22 is a perspective view of a representative embodiment of an
integrated expander/compressor
FIG. 23 is a schematic diagram of a circuit of certain components
of an apparatus used for evaluating the first representative
embodiment of an integrated expander/compressor.
FIG. 24 is a plot, obtained during evaluation of the first
representative embodiment of the integrated expander/compressor, of
isentropic efficiency of the expander operating on compressed
nitrogen (as a working fluid) at room temperature.
FIG. 25 is a plot, obtained during evaluation of the first
representative embodiment of the integrated expander/compressor, of
the transmission efficiency (shaft work relative to work done on
the piston face) of the expander as a function of intake pressure
of nitrogen used as a working fluid.
FIG. 26 is an array of P-V curves, obtained during evaluation of
the first representative embodiment of the integrated
expander/compressor, recorded from one expander cylinder operating
at 1500 rpm shaft velocity.
FIG. 27 is a plot of the isentropic efficiency of the expander, of
the first representative embodiment of an integrated
expander/compressor, operating with HFE-7000 as the working
fluid.
FIG. 28 is an array of P-V curves, obtained during evaluation of
one expander cylinder of the first representative embodiment of the
integrated expander/compressor, operating at a constant inlet
pressure and varying shaft velocity.
FIG. 29 is a plot, obtained during evaluation of the first
representative embodiment of the integrated expander/compressor, of
isentropic efficiency of the compressor operating with
room-temperature air, as a working fluid, drawn through a
pressure-reducing regulator.
FIG. 30 is a plot of transmission efficiency calculated from the
data shown in FIG. 29.
FIG. 31 is an array of P-V curves recorded from one cylinder of the
compressor of the first embodiment of the integrated
expander/compressor.
FIG. 32 is a plot of the isentropic efficiency of the compressor,
of the first embodiment of the integrated expander/compressor,
operating on HFE-7000 as a working fluid and at various stated
shaft rotational speeds.
FIG. 33 is an array of P-V curves, obtained during an evaluation of
one of the cylinders of the compressor evaluated in FIG. 32, having
an inlet vacuum of 6 psi.
FIG. 34 is a schematic diagram of a test circuit used for
performing further evaluations of the first representative
embodiment of the integrated expander/compressor. The depicted test
circuit is configured to perform an energy balance on an operating
expander/compressor.
FIG. 35 is an array of P-V curves, obtained using the test circuit
shown in FIG. 34, of a compressor operating at 500 rpm, with air
being used as the working fluid.
FIG. 36 is a plot of isentropic efficiency of the compressor
evaluated in FIG. 35, operating at a range of shaft speeds and
exhaust-to-inlet pressure ratios.
FIG. 37 is a P-V diagram showing the effect, upon the P-V behavior
of the compressor evaluated in FIG. 35, of using refrigerant as the
working fluid.
FIG. 38 is an array of P-V curves, obtained during evaluation of
the first representative embodiment of the integrated
expander/compressor using the system shown in FIG. 34, recorded
with the expander operating at 1500 rpm while the inlet pressure
varied from 35 psia to 85 psia.
FIG. 39 is a plot, obtained during evaluation of the first
representative embodiment of the integrated expander/compressor
using the system of FIG. 34, of calculated isentropic efficiency of
the expander as a function of shaft speed for various inlet
pressures.
FIG. 40 is a schematic diagram of a representative
microtubule-based configuration that can be used for any of the
several heat-exchanger components of the heat-pump system.
DETAILED DESCRIPTION
Representative Embodiment of Heat-Pump System
A schematic diagram of an embodiment of a heat-pump system 10 (also
termed a heat-pump "cycle") including heat recovery is shown in
FIG. 1. The depicted system 10 represents a combination of a
Rankine power cycle 12 and a vapor-compression refrigeration cycle
14. (A "cycle" in thermodynamic terms is a process in which a
working fluid undergoes a series of state changes and finally
returns to its initial state. A cycle plotted on a diagram of
properties of the working fluid forms a closed curve.)
Unlike a conventional vapor-compression system that requires a
substantial supply of electrical power for the compressor, the
depicted system 10 relies on the work output of an expander 16 to
drive a compressor 18. The system 10 utilizes a working fluid for
heat cycling, and the same working fluid is used throughout the
system.
The power cycle 12 includes a combustor 20, a "boiler" 22, the
expander 16, a "regenerator" 24, a condenser 26, and a liquid pump
28. The combustor 20 generally is a component in which heat energy
is produced (in this embodiment by burning a fuel) for adding to
the system 10 that can be utilized by the system for producing
work, as described later below. The boiler 22 generally is a
component in which the working fluid, passing through the boiler,
is heated (by input to the boiler of heat energy produced by a
suitable source, which in this embodiment is the combustor 20)
sufficiently to vaporize the working fluid and to superheat the
vapor for producing work. The "regenerator" 24 is generally a
heat-exchanger, which recovers heat from where it otherwise would
be wasted and contributes the recovered heat to where it can be
efficiently utilized. The condenser 26 generally is a component in
which the working fluid is converted from a saturated vapor at
elevated pressure to a liquid, which causes the working fluid to
release heat in the condenser.
The refrigeration cycle 14 includes an evaporator 30, the
compressor 18, the condenser 26, and a throttling valve 32 (also
called an "expansion valve"). The evaporator 30 is generally a
component in which working-fluid liquid under reduced pressure is
evaporated, i.e., converted from a liquid into a vapor, which
causes the working fluid to absorb heat in the evaporator. The
compressor 18 generally is a component that increases the pressure
of a working-fluid vapor as the vapor passes through the
compressor. The throttling valve 32 generally is a component that
imposes a substantial flow restriction to passage of working fluid
liquid such that, as working fluid flows through the valve from
upstream (where the fluid is at elevated pressure), the pressure of
the working fluid drops sufficiently so that evaporation can occur
in the downstream evaporator 30.
Fuel is supplied from a fuel tank 34 to the combustor 20, in which
the fuel is combusted in the presence of air, oxygen, or other
oxidizer. Since the same working fluid is used for both cycles 12,
14, a single condenser 26 can be used for both cycles.
It will be appreciated that several components of the system 10
are, in effect, heat-exchangers. Heat-exchanging components
include, for example, the boiler 22, the evaporator 30, and the
condenser 26, as well as the regenerator 24.
In the system 10 shown in FIG. 1, the working fluid (condensed into
a liquid by the condenser 26) is split into two streams. A first
stream 36 enters the refrigeration cycle 14 via the throttling
valve 32, and a second stream 38 enters the power cycle 12 via the
pump 28. In the power cycle 12, as in a Rankine cycle, the
condensed working fluid is delivered at elevated pressure by the
pump 28 through the regenerator 24 to the boiler 22 where the
liquid working fluid is heated (using heat generated in the
combustor 20), desirably at substantially constant pressure, to
form a superheated vapor of the working fluid. The superheated
vapor expands (ideally adiabatically and isentropically) in the
expander 16 that generates a work output ("W.sub.out") from the
expansion. Expansion of the working fluid in the expander 16
decreases the temperature and pressure of the vapor, but the vapor
is still capable of surrendering heat, in the regenerator 24, to
the working fluid being urged by the pump 28 into the boiler 22.
The vapor passes through the regenerator 24 to the condenser 26
where the vapor is converted to a saturated liquid as the working
fluid gives off heat. The saturated liquid reenters the pump 28 and
the cycle repeats.
In the power cycle 12, as the working fluid passes from the pump 28
through the regenerator 24, heat is added to the working fluid
before the working fluid returns to the boiler 22. The added heat
is supplied by the regenerator 24 directly from heat removed by the
regenerator 24 from the vapor exiting the expander 16. Thus, the
regenerator 24 is effectively a heat-exchanger, in which heat
recovery from the expanded working fluid increases the efficiency
with which the working fluid is superheated, thereby providing the
system 10 with a higher overall cycle efficiency compared to
conventional systems. If the regenerator 24 were not included in
the system 10, the excess heat remaining in the expanded working
fluid (exiting the expander 16) would not be recovered, resulting
in reduced performance of the system.
In the depicted system 10, the expanded working fluid (from the
expander 16) exiting the regenerator 24 is combined with the
working fluid of the refrigeration cycle 14, specifically with the
compressed working fluid exiting the compressor 18. The combined
stream of saturated vapor at elevated pressure is routed to the
condenser 26, which converts the vapor working fluid into liquid,
producing heat. (It is possible to provide a second regenerator in
the refrigeration cycle 14 to recover at least some of this heat
from the working fluid and to transfer the recovered heat to
working fluid entering the boiler 22.) Exit of the working fluid
from the condenser 26 completes both the power cycle 12 and the
refrigeration cycle 14. The work output by the expander 16 is input
directly to the compressor 18 to drive the compressor in the
refrigeration cycle 14.
In the refrigeration cycle 14, liquid working fluid from the
condenser 26 passes through the throttling valve 32, which reduces
the pressure of the liquid. The reduced-pressure liquid enters the
evaporator 30, in which the liquid is converted to a vapor, which
substantially reduces the temperature of the working fluid. Thus,
the evaporator 30 achieves net movement of heat to the working
fluid as the pressure of the working fluid is maintained at
saturation conditions by the compressor 18.
The system 10 also includes "balance of plant" components. For
example, the system 10 includes means for driving the pump 28,
means for urging flow of air into the combustor 20, and means for
urging flow of air through the condenser 26. The means for driving
the pump 28 can be any of various devices utilizing a portion of
the work output from the expander 16, for example, a direct-drive
or other coupling to a rotating expander shaft. The respective
means for urging flow of air can be respective fans or the like,
again utilizing respective portions (which would be very small) of
the work output from the expander 16. Balance of plant can include,
if necessary, means for urging flow of fuel from the fuel tank 34
to the combustor 20.
Expansion of the working fluid in the evaporator 30 converts the
working fluid from a reduced-pressure liquid to a reduced-pressure
vapor, which is accompanied by a substantial decrease in
temperature (cooling) of the working fluid. This cooled working
fluid can be used to chill a coolant fluid such as water. Hence,
the system 10 also desirably includes a cycle 40 by which heat is
drawn from the coolant fluid in the evaporator 30 and is circulated
to a "cooling zone" for cooling purposes. In the cycle 40, the
evaporator 300 effectively serves as a heat-exchanger that achieves
transfer of heat from the coolant fluid to the working fluid. The
resulting chilled coolant fluid in the cycle 40 is circulated from
the evaporator 30 to a "cooling zone" 42 of a thing 43 to be
cooled. An example cooling zone 42 is in or on a cooling vest
("water chest") 3v that can be, for example, worn around the chest
of a person 43p in the manner of a part of a suit of clothing. As
the coolant fluid circulates in the cooling zone 42 of the cooling
vest 3v, the coolant fluid removes heat from the body of the person
43p and thus cools the person 43p. The heat thus acquired by the
coolant fluid subsequently is removed by passing the coolant fluid
through the evaporator 30. The cycle 40 can include a small pump
(not shown) used for circulating the coolant fluid through the
cooling vest 3v and evaporator 30. Alternatively to a cooling vest
43v or the like, the cooling zone 42 can be a heat-exchanger used
for cooling air in a vehicle 43c, wherein the air in the vehicle is
circulated so as to flow past the cooling zone.
In the depicted system 10, the expander 16 and compressor 18 are
effectively integrated, and the integrated expander/compressor is
key to achieving overall system efficiency due to, inter alia, less
flow resistance, less heat loss, and greater mechanical efficiency
in the integrated configuration. An exemplary integrated
expander/compressor exhibits an isentropic efficiency of at least
70%. Also effectively integrated are the boiler 22 and combustor
20; in the integrated configuration heat produced in the combustor
is transferred directly and with minimal loss to the working fluid
in the boiler.
At least one (preferably more than one) of the heat-transfer
(heat-exchange) components (e.g., the boiler 22, the evaporator 30,
the condenser 26, and the regenerator 24) desirably is configured
microtechnologically for high thermal efficiency, compactness, and
low mass. Examples of microtechnology-based configurations include,
but are not limited to, thin-walled microfluidic channels and/or
microtubules. Components having such configurations allow the
weight and size of the system 10 to be reduced substantially over a
conventional heat-pump system.
An exemplary microtubule-based configuration 500 that can be used
for any of various heat-exchangers of the system 10 is shown in
FIG. 40, which depicts a housing 502 having a first end 504 and a
second end 506 defining respective end-chambers 508, 510. The first
end-chamber 508 has an inlet port 512, and the second end-chamber
has an outlet port 514. The first end-chamber is bounded by a
barrier 516, and the second end-chamber is bounded by a barrier
518. Microtubules 520 are mounted to and extend between the
barriers 516, 518 such that first ends of the microtubules 520 are
attached to the barrier 516, and second ends of the microtubules
are attached to the barrier 518. The microtubules 520 are attached
to the barriers 516, 518 such that the lumina of the microtubules
extend through the barriers and open into the respective
end-chambers 508, 510. A stream of the working fluid enters the
first chamber 508 via the inlet port 512 and enters (note arrows)
the lumina of the microtubules 520. The working fluid thus flows in
parallel through the microtubules 520 from the first end-chamber
508 to the second end-chamber 510, and exits the second end-chamber
through the outlet port 514. In a middle chamber 524 (also termed a
"shell") of the housing, between the barriers 516, 518, are mounted
baffles 522 arranged in a staggered arrangement. The middle chamber
has an inlet port 526 and an outlet port 528 (depicted for
counter-current flow). A fluid intended to undergo heat-exchange
with the working fluid enters the middle chamber 524 via the inlet
port 526, flows through the middle chamber in a convoluted path
dictated by the baffles 522, and exits via the outlet port 528.
Meanwhile, thermal exchange between the two fluids occurs across
the walls of the microtubules 520.
The microtubule-based configuration 500 shown in FIG. 40 can be
used, for example, as an integrated boiler 22 and combustor 20, in
which hot gases produced by fuel combustion are produced in, or
otherwise flow through, the middle chamber 524 spanned by the
microtubules 520. As working fluid passes through the lumina of the
microtubules 520, the working fluid readily absorbs heat,
transferred quickly across the microtubule walls, from the hot
combustion gases in the middle chamber 524. The microtubules in
such a configuration are made of a suitable metal or ceramic
material capable of withstanding combustion conditions while
providing high thermal conductivity across their walls.
In an alternative microtubule-based configuration of an integrated
boiler/combustor, the hot combustion gases flow through the lumina
of the microtubules 520 as the working fluid flows through the
middle chamber ("shell") 524. This configuration may be
advantageous for certain applications since its comparatively
"open" flow arrangement afforded by the shell allows vaporization
to proceed without flash points that otherwise would tend to "spit"
out both liquid and vapor if occurring in the confines of
small-diameter tubes. This alternative configuration is especially
advantageous in microscale boilers having a microtubule
configuration.
An exemplary microfluidic-channel configuration that can be
utilized (as an alternative to the microtubule-based configuration)
is any of the microchannel configurations now making their debut
for use in actively cooling microprocessor chips. Microchannels can
be formed in a substrate by any of various MEMS or other
microfluidic-fabrication techniques currently available, by which
channels, conduits, through-cuts, and any of various other machine
shapes and voids are formed. The microchannels desirably are
separated from each other by thin walls, and fluids are passed
through opposing microchannels by, e.g., countercurrent flow as
thermal exchange between the fluids occurs across the intervening
walls. Two-dimensional arrays of microchannels can be formed by
existing MEMS technology in laminar substrates, and
three-dimensional arrays of microchannels can be formed by bonding
the laminae together. Examples of applicable microchannel
technology can be found in U.S. Pat. Nos. 6,892,802, 5,932,940,
5,749,226, and 5,811,062, all incorporated herein by reference.
Another exemplary use of microtubules and/or microchannels is in
the regenerator 24. In a regenerator configured with microtubules,
for example, liquid working fluid from the pump 18 can be directed
through the lumina of a bundle of microtubules 520 spanning a
middle chamber 524 traversed by hot working-fluid vapor from the
expander 16. As the liquid working fluid passes through the
microtubules 520, the working fluid readily absorbs heat,
transferred quickly across the microtubule walls, from the
working-fluid vapor that has exited the expander 16.
With respect to the system 10, it is desirable to keep expansion
spaces (e.g., cylinders, pistons, and conduits for the working
fluid) at the temperature of the working fluid during operation. It
also is desirable to minimize, to the best extent possible, heat
leakage to ambient-temperature regions of the system 10 or to the
ambient surroundings of the system 10. Such thermal management can
be achieved by judicious application and use of insulation, such as
around the expander 16 and around other "hot" sections of the
system, and/or use of materials having low thermal conductivity.
Thermal insulation also can be achieved using vacuum-gap
technology.
The system embodiment shown in FIG. 1 has particular utility as a
heat pump (heat-activated cooling system) usable, for example, as a
cooling system for a protective suit worn by a person, a cooling
system for a vehicle, or as a waste-heat-recovery system. The
system 10 has been evaluated both with and without the regenerator
24 situated between the pump 28 and boiler 22. As noted above,
including the regenerator 24 allows effective recovery of heat left
in the expanded working fluid (superheated in the boiler 22)
exiting the expander 16. Herein, a system 10 including the
regenerator 24 is regarded as including a "heat-recovery" cycle. A
system lacking the regenerator 24 is regarded as a "basic cycle."
Thus, the regenerator 24 is a heat-exchanger, i.e., a heat-recovery
component.
Any of various working fluids can be used with the system 10, and a
suitable working fluid for a particular application of the system
will involve considerations of environmental issues, flammability,
toxicity, and the like. The selection can be made from several
general classes of working fluids commonly used in refrigeration. A
first general class is hydrocarbons, including propane (R290),
isobutane (R600a), n-butane (R600), cyclopropane (RC270), ethane
(R170), n-pentane (R601), and isopentane (R601a). A concern with
this first class is the flammability of the compounds; on the other
hand, they have no adverse effect on the earth's ozone layer, are
not generally implicated in global warming, and have low
environmental impacts in production. A second general class is
chlorohydrocarbons (e.g., methyl chloride (R40)). A third general
class is chlorofluorocarbons (e.g., trichlorofluoromethane (R11),
dichlorodifluoromethane (R12), monofluorodichloromethane (R21), and
monochlorodifluoromethane (R22), and trichlorotrifluoroethane
(R113), as well as R114, R500, and R123 (or HCFC-123)). A concern
with the second and third classes is the adverse effect of these
compounds, when released into the environment, on the earth's ozone
layer. A fourth general class is fluorohydrocarbons (e.g.,
tetrafluoroethane (R134a), pentafluoroethane (R125), R502, R407C,
R410, and R417A, and HFE-7000). A fifth general class is other
compounds such as ammonia (R717), sulfur dioxide (R764), and carbon
dioxide. Benefits of the fluorohydrocarbons are their inertness and
non-flammability. Some of these compounds currently have
environmental and/or toxicity concerns associated with them.
Another class of working fluids that may be advantageous for some
uses is nanofluids.
Computational Model of Heat-Pump System
Exemplary operating conditions of the system 10 are set forth in
Table 1.
TABLE-US-00001 TABLE 1 Operational Parameters Value A. Input
Parameter: Evap'r Cooling Load (Q.sub.in) 150 W Evaporator Temp
(T.sub.evap) 7.degree. C. (280 K) Evaporator Pressure (P.sub.evap)
46 kPa Condenser Temp (T.sub.con) 40.degree. C. (313 K) Condenser
Pressure (P.sub.con) 150 kPa Boiler Pressure (P.sub.boi) 1000 kPa
Boiler Temperature (T.sub.boi) 116.degree. C. (389 K) Expander
Isentropic Efficiency (.eta..sub.e) 0.8 Compressor Isentropic
Efficiency (.eta..sub.c) 0.9 Regenerator Effectiveness 0.9 Total
fluid-transfer rate (m) 9.28 .times. 10.sup.-4 kg/s B. Output
Parameter: Heat-pump COP 0.95 Boiler heat output (Q.sub.boi) 159 W
Condenser heat output (Q.sub.con) 309 W Regen'r heat output
(Q.sub.reg) 21 W Expander work output (W.sub.exp) 22 W Pump work
output (W.sub.pump) 0.6 W Boiler mass transfer (m.sub.1) 4 .times.
10.sup.-4 kg/s Evap'r mass transfer (m.sub.7) 5.28 .times.
10.sup.-4 kg/s System weight 1.7 kg
In this example, the cooling load of Q.sub.in=150 Watts and the
evaporator temperature T.sub.evap=7.degree. C. were established by
requirements posed by an effective cooling system for a protective
suit worn by a person in a temperate climate. The temperature of
the condenser T.sub.con was determined by the anticipated
temperature of surrounding air and the required difference of
temperature of the condenser 26 compared to the surrounding air.
Thus, in Table 1, a stated exemplary condenser temperature of
T.sub.con=40.degree. C. was selected.
For high performance in a miniature cooling system, the expander 16
and compressor 18 were integrated together and each provided with a
piston-based configuration. This configuration was suitable for a
moderate-pressure cooling system in which the working fluid has
vapor-pressure characteristics similar to those of isopentane and
various fluorocarbons.
In this computational model the working fluid was isopentane, and
the evaporator 30 and the condenser 26 were regarded as operating
under saturation conditions at respective pressures of
P.sub.evap=46 kPa and P.sub.con=150 kPa. Based on these operating
pressures, the compression ratio required for the compressor 18 was
P.sub.con/P.sub.evap=(150 kPa)/(46 kPa)=3.26. Also, based on the
working-fluid conditions as well as mechanical and material
considerations for the integrated expander/compressor, an exemplary
inlet pressure of the expander 16 was established at P.sub.boi=1000
kPa, which established an expansion ratio of
P.sub.boi/P.sub.evap=(1000 kPa)/(150 kPa)=6.67. At P.sub.boi=1000
kPa and T.sub.boi=116.degree. C., the boiler 22 operates at
saturation conditions. The stated isentropic efficiencies of the
expander 16 and compressor 18 (i.e., .eta..sub.e=0.8 and
.eta..sub.c=0.9) were based on preliminary test results. For a
system 10 including a regenerator 24, the efficiency of the
regenerator was set at 0.9 to recover most of the available heat in
the expanded working fluid exiting the expander 16.
The conditions set forth in Table 1 produced a cycle T-S diagram as
shown in FIG. 2. The T-S diagram included respective portions for
the power cycle 12 and the refrigeration cycle 14. The power cycle
12 comprised the following portions:
path 1-2: isentropic compression of the working fluid in the pump
28
path 2-3: heat addition to the working fluid in the regenerator
24
path 3-4: heat addition to the working fluid in the boiler 22
path 4-5a: expansion of the working fluid in the expander 16
path 5a-6: rejection of heat from the working fluid in the
regenerator 24
path 6-1: rejection of heat from the working fluid in the condenser
26
The refrigeration cycle 14 comprised the following portions:
path 1-8a: throttling the working fluid through the throttling
valve 32
path 8a-9: absorption of heat from the working fluid in the
evaporator 30
path 9-7a: compression of the working fluid in the compressor
18
path 7a-1: rejection of heat from the working fluid in the
condenser 26
The working-fluid vapor exiting the expander 16 remains superheated
even though the regenerator 24 removes some heat from the working
fluid. The respective working-fluid streams from the power cycle 12
and refrigeration cycle 14 combine at 7a in FIG. 2 and then undergo
condensation in the condenser 26, which removes heat from the
fluid. Using isopentane as the working fluid, compression of the
fluid yields some condensation of the fluid. According to the T-S
diagram, a slight superheating of the "refrigerant" working fluid
would avoid such condensation. Slight superheating could be
accomplished by including a regenerator within the refrigeration
cycle 14.
To assess the influence of each component on the overall
performance of the system 10, a thermodynamic model was developed
using EES ("Engineering Equation Solver," available from F-chart
Software, Madison, Wis.) as a data base and equation solver. This
computation package provided all of the thermodynamic and transport
properties needed for the working fluid. To simplify the
thermodynamic model, the following assumptions were made: (1) the
system operated under steady-state conditions; (2) pressure drops
in the boiler 22, regenerator 24, evaporator 30, and condenser 26
and in the connecting conduits were negligible; (3) heat losses
from all the components (except the condenser 26) to the ambient
environment were negligible; (4) the temperature rise across the
liquid pump 28 was negligible; (5) the fluid enthalpy did not
change across the throttling valve 32; and (6) the work output by
the expander 16 equaled the work input to the compressor 18.
In the model, the Martin-Hou equation of state was used with
respect to the working fluid to determine unknown thermodynamic
properties. Based on the operating conditions listed in Table 1 for
the major components in the system 10, two sets of equations were
formulated for the power cycle and the refrigeration cycle. To
complete the overall heat-pump cycle, three additional equations
were used to couple the two cycles together: {dot over
(m)}.sub.tot={dot over (m)}.sub.p+{dot over (m)}.sub.r (1)
P.sub.exp=P.sub.com (2) W.sub.exp=W.sub.com (3) in which:
{dot over (m)}.sub.tot=total mass-flow rate
{dot over (m)}.sub.p=mass-flow rate in the power cycle 12
{dot over (m)}.sub.r=mass-flow rate in the refrigeration cycle
14
P.sub.com=outlet pressure of the compressor 18
P.sub.exp=outlet pressure of the expander 16
W.sub.com=work input to the compressor 18
W.sub.exp=work output by the expander 16
The model yielded calculations of the heat input to, and heat
output from, the system 10 based on the total mass-flow rate {dot
over (m)}.sub.tot. Then, the coefficient of performance (COP) of
the overall system 10 was calculated from Equations (4) and (5),
below. Equation (4) applied to a "basic" cycle lacking the
regenerator 24, and Equation (5) applied to a cycle ("heat-recovery
cycle") including the regenerator.
##EQU00001## in which:
COP.sub.bas=coefficient of performance of the basic cycle
COP.sub.reg=coefficient of performance of the heat-recovery
cycle
Q.sub.boi=heat input to the boiler 22
Q.sub.eva=heat input to the evaporator 30
Q.sub.reg=heat exchanged in the regenerator 24
W.sub.pump=work input to the pump 28
For a particular cooling-system design, the cooling load of (heat
input to) the evaporator 30, Q.sub.eva, was a key parameter that
determined the total mass-flow rate ({dot over (m)}.sub.tot) of the
working fluid for the prescribed operating conditions. This
parameter also determined the needed heat input (Q.sub.boi) in the
boiler 22, which in turn determined the heat rejected in the
condenser 26. As shown in FIG. 1, heat input to the boiler 22 is
supplied by hot gases produced by combustion of fuel. By
configuring the boiler 22 using microtechnology-based structures
(e.g., microchannels and/or microtubules), the boiler can be
provided with high heat-transfer effectiveness. In the following
cycle cases, heat loss via combustion exhaust was not used to
determine the overall system COP.
As noted previously, the basic cycle is similar to that shown in
FIG. 1, but lacks a regenerator 24. In the basic cycle, liquid
working fluid exiting the pump 28 enters the boiler 22 directly
without being preheated. Working-fluid vapor exiting the expander
16 is conducted directly to the condenser 26 without being
subjected to heat recovery. For a cooling load of Q.sub.in=150 W in
the evaporator 30, the total mass-flow rate of working fluid for
the entire cycle ("measured" at the outlet from the condenser 26)
is 0.93 g/s, whereas the mass-flow rates for the power cycle 12 and
refrigeration cycle 14 are 0.4 g/s and 0.53 g/s, respectively. The
heat input to the boiler 22 is 179 W and the heat rejection by the
condenser 26 to the cooling air is 329 W. Therefore, the overall
COP of the heat-pump system configured according to the basic cycle
is 0.83.
An exemplary system 10 providing heat-recovery (due to the presence
and use of the regenerator 24) is shown in FIG. 1. For the same
cooling load of 150 W in the evaporator 30, the mass-flow rates are
the same as in the basic cycle. However, with the effectiveness of
the regenerator 24 being 0.9, heat input to the boiler 22 is
reduced to 159 W, and heat rejection by the condenser 26 is
correspondingly reduced to 309 W. As a result, due to use of the
regenerator 24, the overall COP of the system 10 is increased to
0.96. Although including the regenerator 24 adds system complexity,
the thermodynamic performance of the system is increased
significantly. The regenerator 24 also has additional benefits that
impact respective configurations of the condenser 26 and boiler
22.
Further computations demonstrated that overall system performance
was further improved by using a regenerator 24 whenever the working
fluid is superheated. FIG. 3 shows the effect of the regenerator 24
on the COP of the total system whenever the pressure of the boiler
22 remains at 1000 kPa and superheat is increased from zero (at a
boiler temperature of 116.degree. C.) to 227.degree. C.
To optimize system performance, investigations were made of the
influence of operating conditions for each component. Since the
system 10 can be used advantageously for any of various
portable-cooling applications, overall system weight can be a major
consideration for system design. For example, the lower the
temperature of the condenser 26, the higher the COP of the cycle;
but, reducing its temperature makes the condenser larger (and
heavier) due to a lower .DELTA.T relative to the ambient
environment. A heavier condenser 26 can be disadvantageous for
certain types of portable cooling systems. On the other hand,
increasing the temperature of the condenser 26 to reduce the
condenser weight causes the COP of the refrigeration cycle 14 to
decrease. To consider this effect, and to investigate the
consequences of changing isentropic efficiencies of the expander 16
and compressor 18, the temperature of the boiler 22, the
temperature of the condenser 26, and the temperature of the
evaporator 30, trade-off studies were conducted.
Superheating the working fluid in the boiler 22 can have a
significant impact on system performance. A system lacking a
regenerator 24 has an overall COP that decreases with increasing
superheat of the working fluid because the cooling capacity of the
evaporator 30 increases more slowly than the heating requirement
for the boiler 22. On the other hand, in a system including a
regenerator 24, the overall system COP increases with increasing
superheat of the working fluid because the regenerator 24 recovers
most of the extra heat input to the working fluid in the boiler 22.
These trends are clearly shown in FIG. 3, in which the overall COP
of the system increases almost 40 percent over a 100 K increase of
superheat in the boiler 22 at 1000 kPa. Due to the significant
improvement of performance of a system 10 including a regenerator
24 for heat recovery, the parameters of such a system that included
a combined (integrated) expander/compressor were further evaluated.
A "combined" or "integrated" expander/compressor is a component in
which the expander 16 and compressor 18 are brought together, such
as in a single housing in which certain moving parts of the
expander and compressor are coupled to a single shaft.
As a key component in the system 10, the performance of the
integrated expander/compressor significantly impacts the system
COP. As shown in FIG. 4, the COP of the system 10 decreases almost
linearly with a corresponding decrease in efficiency of the
expander 16. The depicted curve was generated by starting with
baseline values of 80% and 90% isentropic efficiencies for the
expander 16 and compressor 18, respectively, and by decreasing the
efficiency of the expander while keeping the compressor efficiency
equal to expander efficiency plus 0.1.
Increasing the fluid-condensing temperature dramatically decreases
the mass of the condenser 26 due to the increase of the log mean
temperature difference (LMTD) of the condenser. But, this decrease
in the mass of the condenser 26 is accompanied by a corresponding
decrease in system COP. In a trade-off study of condenser
temperature versus system weight, optimal operating temperatures
for the condenser 26 (taking into consideration system weight and
performance) were determined. In the trade-off study, the system
weight is the total weight of the overall system 10. System weight
includes the respective masses of all the components shown in FIG.
1 plus a cooling fan (not shown) for the condenser 26. FIGS. 5 and
6 show the effects on system weight and COP, respectively,
accompanying increases in temperature of the condenser 26.
An increase in size of the evaporator 30 accompanies an increase in
the evaporator temperature. Under such conditions, both the heat of
vaporization and the fluid mass-flow rate in the refrigeration
cycle 14 increase, yielding substantial increases in the cooling
capacity of the evaporator 30. FIG. 7 shows the influence of
evaporator temperature on the overall system COP. For example, a
5-degree increase in the evaporation temperature yields an increase
of 20 percent in the overall system COP.
Thus, the thermodynamic model revealed that a system 10 including
heat recovery (using at least one regenerator 24) exhibited
significantly better performance than a system lacking a
regenerator, especially whenever the working fluid in the boiler 22
is superheated. Overall system COP increased almost 40 percent with
100 degrees of superheat in the boiler 22. A further improvement of
20% could be realized if the evaporator temperature were increased
from, for example, 7.degree. C. to 12.degree. C. Overall system COP
dropped approximately linearly with corresponding decreases in the
respective efficiencies of the expander 16 and compressor 18.
Although overall system COP also dropped linearly with increasing
condenser temperature, the accompanying effect of reducing overall
system size and weight was regarded as beneficial for certain uses
of the system.
First Representative Embodiment of Expander
The expander embodiment described below is advantageously used in
small, compact heat-pump systems such as a system used for cooling
a personal protective suit.
The expander of this embodiment converts thermal energy, added to
the working fluid in the boiler, into shaft work used for driving
the compressor and optionally other components of the heat-pump
system. In this embodiment, the expander has a piston
configuration, in which the pistons are disposed in respective
"cylinders." Each cylinder includes a respective inlet valve and a
respective "exhaust" (outlet) valve.
With respect to the expander, FIGS. 8(a)-8(c) schematically show an
exemplary piston disposed in its respective cylinder 52. The
cylinder 52 includes a cylinder head 54 in which an inlet valve 56
and an exhaust valve 58 are mounted. FIGS. 8(a)-8(c) depict three
respective steps of a piston cycle. The piston cycle includes one
downward "stroke" and one upward "stroke" of the piston 50 in the
cylinder 52. FIG. 8(a) depicts the start of a downward stroke, in
which the piston 50 is at top-dead-center (TDC), the exhaust valve
58 is closed, and the inlet valve 56 is open, which allows
working-fluid vapor to enter the cylinder 52. This intake of vapor
into the cylinder 52 is isobaric. In FIG. 8(b), when the piston 50
has traveled about 30% of the downward stroke, the inlet valve 56
closes, which initiates an expansion of the vapor in the cylinder
52. Expansion continues with further downward movement of the
piston 50 to its bottom-dead-center (BDC) position in the cylinder
52 (FIG. 8(c)). At BDC, the exhaust valve 56 opens to discharge the
expanded vapor from the cylinder 52 as the piston 50 returns via an
upward stroke to the TDC position. To avoid flow-through in the
cylinder 52 from the inlet valve 56 directly to the exhaust valve
58, the exhaust valve can be closed just before the piston 50
reaches TDC, which causes a slight compression of vapor remaining
in the cylinder. (Technically, this would constitute a fourth and
final step before TDC, but the impact on the cycle work would be
small.)
Turning to FIG. 9(a), this expander embodiment 60 comprises two
pistons 50a, 50b each situated in a respective cylinder 52a, 52b.
By way of example, each piston has a diameter of 0.5 inches and a
stroke of 0.48 inches, yielding a total displacement of
2.pi.(0.48)[(0.5)/2].sup.2=0.189 in.sup.3 (for both cylinders) in
the expander 60. Each piston 50a, 50b is coupled by a respective
rod 62a, 62b and bearing 64a, 64b to a shaft 66. Each cylinder 52a,
52b includes a respective inlet valve 56a, 56b and respective
exhaust valve 58a, 58b. Fixed valve timing provides a
volume-expansion ratio of 3.1 in each cylinder 52a, 52b. Movement
of the pistons 50a, 50b in the respective cylinders 52a, 52b is
180.degree. out of phase (i.e., as one piston undergoes a downward
stroke, the other piston undergoes an upward stroke; when one
piston is at TDC, the other piston is at BDC).
The pistons 50a, 50b are coupled to the shaft 66 by a kinematic
linkage, as noted above, comprising connecting rods 62a, 62b and
bearings 64a, 64b. The bearings are mounted on respective eccentric
disks on the shaft 66. The bearings and eccentric disks reduce
frictional effects during rotation of the shaft 66. If significant
side loads are present in the motion of the pistons, low-friction
material can be used on the sides of the pistons to reduce
parasitic effects during operation.
FIG. 9(b) depicts further detail of the upper portion of a piston
50, cylinder 52, cylinder head 54, inlet valve 56, and exhaust
valve 58 of the expander 60. The cylinder 52 and piston 50 can be
made of any suitable rigid material capable of withstanding
exposure to the working fluid under the temperature and pressure
conditions encountered in the expander 60. In this embodiment the
cylinder 52 and piston 50 are made of an aluminum bronze for
wear-resistance and low friction. The cylinder head 54 in this
embodiment is made of a 0.090-inch thick aluminum plate. Between
the cylinder head 54 and cylinder 52 is an elastomeric seal
(O-ring) 55. Mounted to the cylinder head 54 is a valve block 68
made, e.g., of a rigid polymer such as Delrin or of metal. The
valve block 68 defines an inlet passage 70 and an exhaust passage
72 extending to the cylinder head 54. Situated between the cylinder
head 54 and valve block 68 is an elastomeric (e.g., silicone)
gasket 73 that forms a seal between the cylinder head and valve
block and serves as a seat material for the valves 56, 58. Attached
to the top of the piston 50 is a cap 74 that houses two
spring-loaded tappets 76 (for the inlet valve 56), 78 (for the
exhaust valve 58). Note respective springs 80, 82. The cap 74 also
defines a running surface 83 for the piston 50 in the cylinder
52.
The inlet valve 56 and exhaust valve 58 include respective
spring-loaded poppets 84, 86 that are situated and configured to
engage the respective tappets 76, 78. Engagement of a tappet 76, 78
with a respective poppet 84, 86 actuates the respective valve 56,
58. Thus, actuation of the valves 56, 58 relies upon and is
synchronized with the motion of the piston 50. Because of its
simplicity and inherent valve-timing characteristic, this
configuration is effectively used in miniaturized expanders.
In this embodiment the inlet tappet 76 first contacts the inlet
poppet 84 on the return stroke when the piston 50 is at about 30%
of the stroke from top-dead-center (TDC). The spring 80 associated
with the inlet tappet 76 is selected such that its spring constant,
upon compression, cannot overcome the pressure difference across
the closed inlet valve 56. This allows the tappet spring 80 to
compress while the inlet valve 56 remains closed. Just before TDC,
the inlet tappet 76 encounters a hard-stop that prevents further
compression of the spring 80 and pops the inlet valve 56 open. With
the pressure difference across the inlet valve 56 thus relieved,
the tappet spring 80 pushes the inlet valve 56 fully open and holds
it open until the piston 50 has moved down the downward stroke
sufficiently to fully extend the inlet tappet 76. As the piston 50
moves further on the downward stroke, a spring 88 associated with
the inlet poppet 84 (this spring 88 is "softer" than the tappet
spring 80) closes the inlet valve 56.
The exhaust valve 58 is actuated by the exhaust tappet 78 that
pushes the exhaust-valve poppet 86 to its closed position just
before the inlet tappet 76 reaches its hard-stop. The spring 82
associated with the exhaust tappet 78 absorbs the small amount of
interference between the exhaust tappet 78 and the exhaust poppet
86. A rocker arm 90 (the end of the rocker arm 90 is shown)
contacts the top end of the exhaust poppet 86 of the depicted
cylinder 52 as well as the top end of the exhaust poppet of the
second cylinder (not shown) of the expander. Thus, the rocker arm
90 couples the respective exhaust valves 58 to each other in a
manner ensuring their actuation 180.degree. out of phase with each
other during operation of the expander; when one exhaust valve 58
is closed, the other exhaust valve is open.
As shown in FIG. 10, the cap 74 of each piston 50a, 50b of the
expander 60 defines a seal 92 comprising a flexible lip 94. The tip
96 of the lip 94 engages the inner surface 98 of the cylinder 52 as
urged by gas pressure on the back side 100 of the lip. I.e.,
whenever the cylinder 52 is pressurized, the tip 96 of the lip 94
is forced radially outward and against the surface 98. Thus,
leakage of gas pressure across the piston 50 is substantially
minimized. The seal 92 (and optionally the inner surface 98)
desirably is made of a soft and "slippery" material such as a
fluoropolymeric elastomer, thereby allowing the expander to be
operated without lubrication. In FIG. 10, the clearance between the
piston 50 and cylinder 52 has been exaggerated; typically, the
outside diameter of the piston is approximately 100 micrometers
less than the inside diameter of the cylinder.
Not intending to be limiting, a Scotch yoke is used in this
embodiment as the kinematic linkage between the piston 50 and the
shaft 66 in the expander 60. The yoke is guided by the piston 50 in
the cylinder 52 on one end and another piston in a cylinder on the
opposite side of the shaft 66. The bearing 54 coupling the piston
rod 62 to the shaft 66 rides in a horizontal slot (serving as a
crank arm) in the yoke moving the piston 50 up and down as the
shaft 66 rotates. The location (x) of the piston 50 relative to TDC
is defined by: x=R(1-cos(.theta.) (6) in which:
R=radius of crank arm
.theta.=rotational position of shaft (TDC=0)
The captured volume (V.sub.cyl) of the cylinder 52 is obtained by
multiplying the distance (1-cos(.theta.)) by the cross-sectional
area of the piston 50 and adding the clearance volume (V.sub.clear)
(i.e., dead space at top of cylinder):
.pi..times..times..function..function..theta. ##EQU00002## in which
D=cylinder diameter. This relationship is used later to obtain
pressure-volume curves from paired measurements of cylinder
pressure and shaft-rotational position.
FIG. 11 shows a basic thermodynamic model of this embodiment. The
working fluid 110 enters the expander 112 at the left and exits the
expander with reduced enthalpy at the right. In addition, there may
be some heat transfer into the expander 112. According to the first
law of thermodynamics, the work produced by the expander 112 must
equal the change in enthalpy of the working fluid 110 flowing
through the expander plus the heat added. By definition, an
isentropic process is one in which there is no heat transfer.
Therefore, temporarily assuming an adiabatic process yields: {dot
over (W)}={dot over (m)}(h.sub.i-h.sub.e) (8) in which:
h.sub.i=specific enthalpy of the inlet mass of the working fluid
110
h.sub.e=specific enthalpy of the working fluid exiting the expander
112
m=mass passing through the expander 112
Maximum work occurs when the process is reversible. In this case,
the exiting enthalpy is at a minimum and is denoted by the
subscript "s". The ratio of the actual work to this isentropic work
is the isentropic efficiency (.eta..sub.s):
.eta..function. ##EQU00003## in which:
h.sub.e,s=specific enthalpy of exiting mass under isentropic
expander operation
W.sub.s=isentropic work potential
If the working fluid is an ideal gas, then its enthalpy is directly
proportional to its temperature. The exit temperature is then a
function of the inlet temperature and the exhaust-to-inlet pressure
ratio, yielding:
.eta..times..function. ##EQU00004## in which:
C.sub.p=constant-pressure specific heat of the working fluid
T.sub.i=temperature of inlet mass
T.sub.e,s is the temperature of the exit mass under isentropic
expander operation, and is expressed as:
.function. ##EQU00005## in which:
k=ratio of constant-pressure to constant-volume specific heats for
working fluid
P.sub.e is exit pressure
P.sub.i is inlet pressure
If the heat transfer into the expander 112 is not zero, the process
is not isentropic, and the maximum possible work can potentially
increase. The efficiency (.eta..sub.p) of this polytropic process
is defined by:
.eta..times..function. ##EQU00006## in which heat transfer is
expressed as: {dot over (Q)}={dot over
(m)}C.sub.p(T.sub.e-T.sub.e,s) (13) in which T.sub.e is the
temperature of the exit mass. I.e., the heat transfer is determined
by the difference in exhaust temperature from the isentropic case.
This permits a way of assessing the degree of heat transfer
occurring during the gas-expansion process.
Note that the presence of heat transfer is not necessarily
detrimental. Whenever heat transfer is sufficient to maintain the
working fluid at constant temperature, the expansion process
approaches the ideal isothermal expansion of the Carnot cycle. The
extent to which heat transfer represents irreversibility depends on
the temperature difference across which the heat is transferred. In
the case of the Carnot expansion process, the temperature
difference is assumed to be zero so that the process is fully
reversible. As a general rule, however, heat transfer leads to
irreversibility.
FIG. 12 shows a more detailed thermodynamic model of the expander
112. The model provides the expander 112 with three portions: the
inlet passage 70, the cylinder 52, and the exhaust passage 72. Heat
transfer (Q.sub.i) to the gas in the inlet passage 70 is assumed to
be negative, and heat transfer (Q.sub.e) to the gas in the exhaust
passage 72 is assumed to be positive. (Q.sub.c is heat transfer to
the gas in the cylinder 52.) Since no useful work is produced by
heat transfer in either of the inlet passage 70 or exhaust passage
72, irreversibility is represented. In fact, the depicted
configuration can be considered a short-circuit path for the
enthalpy of the inlet stream of working fluid to reach the exhaust
stream. However, the processes occurring in these portions are
ideally constant-temperature processes and can be minimized by
using a low-thermal-conductivity material for the valve block
68.
While heat transfer inside the cylinder 52 can produce useful work,
as mentioned earlier, most of the heat transfer is irreversible.
The rapid cycling of gas inside the cylinder 52 causes the cylinder
walls 98 of the cylinder to act as a regenerator. Heat is absorbed
by the walls 98 from the working-fluid vapor during fluid intake
and returned to the vapor during the expulsion of the vapor. As
such, the cyclic heat transfer can be a short-circuit path, for
heat through the expander, that produces no useful work. An
indication of the magnitude of this heat transfer can be obtained
by using pressure-volume data to determine the polytropic exponent
of the expansion phase of the cycle. During this phase, the mass in
the cylinder 52 is fixed and the heat transfer to the vapor
(Q.sub.p) in the polytropic process can be calculated from the
following equation:
.times..times..times. ##EQU00007## in which:
P.sub.1,P.sub.2=initial and final pressures in expansion process,
respectively
V.sub.1,V.sub.2=initial and final volumes in expansion process,
respectively
Note that the ratio on the right is the expression for isentropic
work of an ideal gas. Thus, the ratio of heat transfer to work is
proportional to the difference of the polytropic exponent from the
ratio of specific heats.
Heat transfer in the expander 112 is strongly influenced by the
particular selection of working fluid. Of primary consideration is
the heat-capacity ratio, k.
As k approaches unity, operation of the expander approaches
isothermal operation, as indicated by Equation (11). Reduced
temperature swing in the cylinder reduces the cyclic heat transfer
in the cylinder 52 proportionally. As long as this reduced heat
transfer is not at the expense of the isentropic work potential of
the working fluid (the enthalpy change for a given ratio of inlet
and exit pressures, assuming isentropic expansion) then the losses
in the expander 112 related to heat transfer should be reduced by
lowering k.
Table 2 compares the properties of nitrogen and isopentane as
exemplary working fluids. The numbers are based on an inlet
temperature (T.sub.i) of 130.degree. C. and a pressure ratio of
5:1. The volumetric enthalpy change is based on the change in
specific enthalpy of the gas under isentropic expansion divided by
the final specific volume of the gas. Thus, this parameter
represents a measure of the isentropic work potential for a given
cylinder volume. Table 2 shows that, although slightly less work
per stroke would be obtained with isopentane, the temperature swing
in the cylinder is dramatically less. Accordingly, heat-transfer
losses are potentially much less.
TABLE-US-00002 TABLE 2 Comparison of fluid properties related to
cyclic heat transfer Nitrogen Isopentane Units Heat-capacity Ratio
1.4 1.07-1.11 -- Isentropic Temperature Swing 155 41 .degree. C.
Volumetric Enthalpy Change (h.sub.2 - h.sub.1)/v.sub.2 410 320
kJ/m.sup.3
A dynamometer was constructed to measure the torque produced by the
expander 60 at controlled speeds. The measurement setup included a
20-Watt motor for starting and loading the expander 60 and a
torque/speed sensor for measuring the output power of the expander.
The motor was a brushed DC motor (Maxon type S2322, Sachseln,
Switzerland) with an attached digital encoder for control feedback
to a motor controller (Maxon model 4-Q-DC LSC 30/2, Sachseln,
Switzerland). To allow the motor to operate as a load, a
four-quadrant speed controller was used, and a resistive load was
connected in parallel to the power input of the controller. When
braking, the power produced by the motor offset the power going to
the fixed resistive load. Since the resistor desirably is sized to
absorb the maximum braking load, the power supply desirably is able
to supply this amount plus the maximum driving load.
The torque sensor (dynamometer) was a Model E-300 rotary
non-contacting sensor unit manufactured by Sensor Technology (Upper
Heyford, Bicester, Oxon, UK). The torque sensor measures torque by
measuring the propagation speed of surface acoustic waves induced
on the rotating shaft. The wave-velocity changes in proportion to
stress in the material surface. The range of the torque sensor was
.+-.100 mN-m with a specified accuracy of 0.25% of full scale.
However, transverse loads on the sensor shaft caused by the
couplers in the system reduced the precision of the device to about
1% of full scale.
The process-flow measurement setup is shown in FIG. 13, showing the
expander 60 and torque sensor (dynamometer) 118. For convenience,
dry nitrogen gas was used as the working fluid in performance
testing of the expander 60. The nitrogen gas was supplied from a
tank 120 at room temperature and at 20 psig to 60 psig (regulated
by the pressure regulator 122) to the inlet 124 of the expander 60.
To avoid difficulties in measuring the pulsatile flow of gas to the
inlet 124, flow of gas through the exhaust 126 was measured by
mounting the expander 60 and dynamometer 118 inside an air-tight
box 128 and connecting a bubble flow meter 130 to the exhaust 132
of the box. The dynamometer 118 was mounted inside the box 128 to
avoid drag that otherwise would be caused by having to use rotary
seals. By connecting the exhaust 126 of the expander 60 to a
separate exhaust port (normally capped), the bubble flow meter 130
can also be used to measure the leakage rate across the piston seal
during operation of the expander 60.
Pressure-volume curves of the expander cycle were obtained by using
a piezo-resistive pressure transducer 136 (Endevco Model 8530-50,
San Juan Capistrano, Calif.) to measure pressure inside the
cylinder 52. Data from the transducer were used to generate P-V
curves using a setup as shown in FIG. 14. A digital encoder (not
shown) was coupled to the shaft 66 to allow measurements of
rotational position of the shaft. The signals from the encoder were
converted to corresponding analog signals using a digital-to-analog
converter ("DAC") 138. The signals from the DAC 138 and from the
transducer 136 were routed to and recorded on an oscilloscope 140
and then transferred to a computer 142 that converted the
shaft-angle data to corresponding cylinder-volume data using
Equation (7). The data were exported to a spreadsheet for
analysis.
FIG. 15 is a plot of shaft power as a function of shaft rotational
velocity for various inlet pressures to the expander, and FIG. 16
shows corresponding torque data. The graphs show that torque is
relatively independent of shaft speed, resulting in the linear
power data. The small drop in torque exhibited in the range of
shaft speed is primarily due to increasing pressure drop across the
inlet valve with increases in shaft speed. This can be alleviated
by increasing the diameter of the inlet port. The inlet valves were
determined to close later, relative to shaft rotational position,
above 1500 rpm, which allowed additional gas to flow into the
cylinder. This additional gas resulted in a power boost that offset
the increasing inlet loss.
FIG. 17 is a plot of calculated isentropic efficiency of the
expander as a function of shaft rotational speed for inlet
pressures ranging from 35 psia to 75 psia (20 psig to 60 psig). For
the higher pressures of 55 psia and above, the isentropic
efficiency is consistently in the range of 70% to 80%. At lower
pressures, overexpansion of the gas appeared to generate negative
cylinder pressures before the exhaust valve opened.
Observed decreased efficiencies at lower shaft velocities and
higher pressures may have been caused by slight leaks around
certain seals such as the piston lip seal and head gasket,
especially at higher inlet pressures. Preventing such leaks would
follow routine optimization of sealing surfaces and sealing
materials.
FIG. 18 is a plot of polytropic efficiency of the expander as a
function of shaft velocity for inlet pressures ranging from 35 psia
to 75 psia. The heat transfer was calculated using Equation (13).
The temperature of vapor exiting the expander through the exhaust
valve was measured by inserting a thin-gauge thermocouple into the
exhaust port of one of the cylinders. Because the temperature
reading was very sensitive to thermocouple placement, a significant
portion of the heating likely occurred in the exhaust port. But,
not all the heat transfer is attributable to the exhaust ports. The
polytropic exponents obtained by curve fitting to the expansion
portion of the PV data are consistently about 1.1 at 500 RPM and
1.2 at 2500 RPM. These data indicate that the temperature swing of
the gas during the expansion phase is 25% to 50% of that for
isentropic expansion.
Cooling of the intake gas, either in the inlet passage or in the
cylinder, is indicated by higher mass-flow rates through the
cylinder than predicted by an adiabatic-expander model. Typically,
excess flow was 20% higher than explainable by the measured leak
rates. Some portion of this excess is also due to heating during
the expansion and exhaust phases. Higher exhaust temperature
results in less mass in the cylinder when the exhaust valve
closes.
FIG. 19 is a plot of P-V (pressure-volume) data recorded with the
expander operating at 1500 rpm shaft speed and at an inlet pressure
ranging from 35 psia to 75 psia. Each plot starts at the upper left
and progresses clockwise in a loop. The intake phases are first,
and are indicated by the relatively horizontal portions extending
across the top of each loop. At about 0.03 in.sup.3, the inlet
valve closes and expansion begins. The opening of the exhaust valve
at the end of the expansion is indicated by the abrupt drop in
pressure exhibited by the 75 and 65 psia plots near 0.10 in.sup.3.
The respective exhaust strokes correspond to the substantially
horizontal line extending across the bottom to where the exhaust
valve closes at 0.015 in.sup.3. A slight recompression occurs
before the inlet valve opens just before TDC.
The pressure drop during the intake phase is much more noticeable
than the backpressure during the exhaust phase. It is believed this
discrepancy arises from the difference in the length of conduits
leading to the respective valves. For example, the intake valve was
connected by several inches of conduit to a surge tank at which
supply pressure was measured. In contrast, the exhaust valve vents
directly to the atmosphere.
Representative Embodiment of Compressor
Turning to FIG. 20(a), this compressor embodiment 200 comprises two
pistons 202a, 202b each situated in a respective cylinder 204a,
204b. By way of example, each piston 202a, 202b has a diameter of
1.0 inch and a stroke of 0.48 inches, yielding a total displacement
of 2.pi.(0.48)[(1.0)/2].sup.2=0.754 in.sup.3 (total for both
cylinders) in the compressor 200. Each piston 202a, 202b is coupled
by a respective rod 206a, 206b and bearing 208a, 208b to a shaft
209. Each cylinder 204a, 204b includes a respective inlet valve
210a, 210b and respective exhaust valve 212a, 212b. Movement of the
pistons 202a, 202b in the respective cylinders 204a, 204b is
180.degree. out of phase (i.e., as one piston undergoes a downward
stroke, the other piston undergoes an upward stroke; when one
piston is at TDC, the other piston is at BDC).
The exploded view of FIG. 20(b) depicts details of the cylinders
204a, 204b and valves. The cylinders 204a, 204b are shown, each
with a respective flange 214a, 214b. The flanges 214a, 214b mount
to a valve block 216, with valve components being situated
therebetween. For each cylinder 204a, 204b, the valve block 216
defines a respective inlet port 218a, 218b and a respective exhaust
port 220a, 220b. Passage through the inlet ports 218a, 218b to
inside the cylinders is governed by respective inlet-valve flappers
222a, 222b, and exhaust from the cylinders is governed by
respective exhaust-valve flappers 224a, 224b. The inlet-valve
flappers 222a, 222b are defined in respective inlet-flapper members
225a, 225b that nest in respective voids 226a, 226b defined in
respective inlet-valve spacers 228a, 228b. Each inlet-valve flapper
222a, 222b, when in a closed position, seals against a respective
gasket 230a, 230b of a respective flapper seat 232a, 232b. The
exhaust-valve flappers 224a, 224b are defined in respective
exhaust-flapper members 234a, 234b that nest in respective voids
236a, 236b defined in respective exhaust-valve spacers 238a, 238b.
Each exhaust-valve flapper 224a, 224b, when in a closed position,
seals against a respective gasket 240a, 240b of a respective
flapper seat 242a, 242b. Gaskets 244a, 244b, 246a, 246b complete
the sealing to the valve block 216.
The cylinders 204a, 204b and pistons 202a, 202b can be made of any
suitable rigid material capable of withstanding exposure to the
working fluid under the temperature and pressure conditions
encountered in the compressor 200. In this embodiment the cylinders
204a, 204b and pistons 202a, 202b are made of an aluminum bronze
for wear-resistance and low friction. The valve block 216 in this
embodiment is made of aluminum alloy.
In one embodiment the flapper members 225a, 225b, 234a, 234b are
made from thin sheets (shim stock, 0.002-inch thick) of stainless
steel. The flapper seats 232a, 232b 242a, 242b are made of a
nitrile elastomer covering silicone rubber, as are the gaskets
230a, 230b, 240a, 240b. The gaskets 244a, 244b, 246a, 246b are
respective elastomeric O-rings. In another embodiment the flapper
members were made of 0.003-inch thick spring steel shim.
The inlet valves 210a, 210b and exhaust valves 212a, 212b, as
flapper valves, operate on respective pressure differentials that
exist or are established across each valve. The flapper valves are
passive and require no mechanical actuation. As a piston 202a moves
downward from TDC, a small volume of trapped gas in the cylinder
204a is expanded until the pressure of the gas drops below the
pressure in the conduit leading to the inlet valve 210a, at which
moment the higher pressure in the conduit urges the inlet-valve
flapper 222a open. Working fluid then flows through the open valve
210a into the cylinder 204a until the piston 202a reaches BDC. At
BDC, as the piston 202a reverses stroke direction, the inlet-valve
flapper 222a closes by spring action of the flapper itself. The
fluid in the cylinder 204a is then compressed until the pressure in
the cylinder rises above the exhaust pressure and thus pushes the
flapper 224a of the exhaust valve 212a open. As the piston 202a
continues to move to TDC, the compressed vapor exits the cylinder
204a through the exhaust valve 212a.
The shaft 209 to which the pistons 202a, 202b of the compressor 200
are coupled is the same shaft that is coupled to the pistons of the
expander. Consequently, rotation of the shaft caused by
reciprocation of the pistons of the expander causes corresponding
reciprocation of the pistons 202a, 202b of the compressor 200. In
other words, the compressor 200 is directly linked to the expander
such that operation of the expander directly causes operation of
the compressor, and the compressor exploits at least a portion of
the work generated by the expander. This direct coupling of the
compressor to the expander reduces coupling losses that otherwise
would arise between the expander and compressor if these two
components were not coupled in this manner.
Second Representative Embodiment of Expander
As in the first representative embodiment, the expander of the
second representative embodiment uses a piston-actuated valve
system to control fluid flow into and out of the cylinders. The
general configuration of the cylinders and valves of this
embodiment are as shown in FIGS. 8(a)-8(c) and 9(a). Regarding more
specific details, an elevational section of the upper portion of
one of the cylinders 302 of the expander 300 is shown in FIG. 21.
Also shown is the piston 304 situated within the cylinder 302, and
a portion of the connecting rod 306 coupling the piston 304 to the
shaft (not shown). The piston 304 includes a cap 308 that defines,
inter alia, a seal with respect to the inside surface 310 of the
cylinder (seal not detailed, but see FIG. 10). The cylinder 302 is
mounted to a cylinder head 312 that defines an inlet port 314 and
an exhaust port 316. Mounted to the cylinder head 312 are an inlet
valve 318 and an exhaust valve 320. The inlet valve 318 controls
flow of working fluid from the inlet port 314 into the cylinder
302, and the exhaust valve 320 controls flow of working fluid from
the cylinder out through the exhaust port 316. The inlet valve 318
and exhaust valve 320 extend through respective bores 322, 324
defined in the cylinder head 312.
The inlet valve 318 is a spring-loaded poppet valve comprising a
tappet 326 and associated spring 328 situated in the bore 322. The
tappet 326, in turn, interacts with an inlet-valve poppet 330 that
is spring loaded by a respective spring 332. In operation, the
piston 304 first contacts the inlet tappet 326 about 0.080 inch
before TDC. The tappet spring 328 is configured such that its
spring force, upon compression, cannot overcome the force produced
by the pressure difference across the closed inlet valve 318. This
causes the tappet spring 328 to compress while the inlet valve 318
remains closed. Just before TDC, the piston 304 contacts a nipple
334 extending from the inlet poppet 330, and the resulting force
applied to the inlet poppet forces the inlet valve 318 open.
Opening of the inlet valve 318 relieves the pressure difference
across the inlet valve, allowing the tappet spring 328 to push the
inlet poppet 330 fully open and to hold it fully open until the
piston 304 has moved back down sufficiently so that the inlet
tappet 326 is fully extended. As the piston 304 continues to move
downward, the inlet-poppet spring 332 (which is a "softer" spring
than the inlet-tappet spring 328) pushes the inlet valve 318 closed
again.
The exhaust valve 320 comprises a rod 336 and valve disk 338. The
exhaust valve 320 is actuated by an elastomeric bumper 340 mounted
in the piston cap 308. The bumper 340 protrudes about 0.010 inch
from the upper surface of the piston cap 308 so that the bumper can
contact the disk 338 and thus push the exhaust valve 320 closed
just before the piston 304 contacts the nipple 334 on the inlet
poppet 330. As the exhaust valve 320 closes, it actuates a rocker
arm 342 that couples the respective exhaust valves 320 of the two
cylinders 302 of the expander 300. Because the pistons 304 operate
180.degree. out of phase with each other, the rocker arm 342 causes
the exhaust valve 320 of one cylinder 302 to open at BDC as the
exhaust valve 320 of the other cylinder 302 closes at TDC.
To avoid contamination of other portions of the system, it is
desirable that the expander 300 operate without requiring
lubrication. To such end, the piston caps 308 can be made, for
example, of polyphenylene sulfide (PPS), which provides a
wear-resistant, low-friction running surface for the pistons 304
relative to the inside surfaces 310 of the cylinders 302. The
piston cap is formed with a lip seal similar to that shown in FIG.
10 and discussed in the description of the first representative
embodiment. By way of example, the lip is 0.005-inch thick.
Whenever the cylinder 302 is pressurized, the lip flexes radially
outward slightly to seal against the inside surface 310 of the
cylinder. The piston 304 is coupled by a rod 344 to the shaft (not
shown).
Representative Embodiment of Integrated Expander/Compressor
An integrated expander/compressor 350 according to this embodiment
is shown in FIG. 22 and comprises a pair of expander cylinders
352a, 352b (and respective pistons 353a, 353b located inside the
cylinders) and a pair of compressor cylinders 354a, 354b (and
respective pistons 355a, 355b located inside the cylinders) mounted
to a common crankshaft 356. The expander cylinders 352a, 352b are
mounted to a cylinder head 358 to which the expander valves (not
detailed) are mounted. Similarly, the compressor cylinders 354a,
354b are mounted to a cylinder head 360 to which the compressor
valves (not detailed) are mounted. Exemplary valves for the
expander cylinders 352a, 352b can be as described in the
representative embodiments discussed above (i.e., spring-loaded
poppet valves actuated by spring-loaded tappets mounted on the
expander pistons). Similarly, exemplary valves for the compressor
cylinders 354a, 354b can be as described in the representative
embodiment discussed above (i.e., passive "flapper" valves). In the
embodiment shown in FIG. 22, the expander cylinders 352a, 352b are
above and the compressor cylinders 354a, 354b are below the shaft
356. Further with respect to the figure, the expander inlet 362 is
situated at top left and the compressor inlet 364 is situated at
the bottom center. The expander exhaust ports 366a, 366b are
defined in the cylinder head 358, and the compressor exhaust ports
(defined in the cylinder head 360) are not visible in the
drawing.
In the integrated expander/compressor, the expander portion is a
relatively high-pressure portion, in which the spring-loaded and
-actuated poppet valves (as described above) are advantageously
employed. The compressor portion, on the other hand, is a
relatively low-pressure portion, in which the passive "flapper"
valves, as described above, are advantageously employed. In
addition, the pistons of the expander portion are substantially
non-compliantly linked, via their respective rods and bearings
coupled to the shaft, to the pistons of the compressor.
For evaluations thereof, the integrated expander/compressor was
configured to exhaust into ambient atmosphere. Consequently, the
device was evaluated in a sealed chamber. A low-pressure working
fluid (i.e., a fluid with a vapor pressure less than 1 atmosphere
at room temperature) was selected. The chamber was made of glass to
allow visual inspection of the device during operation.
The expander/compression device was sized to provide up to 150 W of
cooling at 2500 rpm with HFE-7000 working fluid. The device had a
cylinder bore of 0.5 inch for the expander and 1 inch for the
compressor. Both pairs of pistons had a stroke of 0.48 inch,
yielding a total displacement of 0.189 in.sup.3 for both expander
cylinders and 0.756 in.sup.3 for both compressor cylinders.
At this miniature scale, heat transfer can be significant, and its
impact on device efficiency measurements was considered. In the
expander, adding heat during expansion reduces the rate at which
pressure falls as the gas is expanded volumetrically. As a result,
more work can be produced from a given mass drawn into the
cylinder. Hence, an expander operating with heat addition
theoretically can operate with an isentropic efficiency greater
than 100%. Conversely, removing heat from the compressor also
reduces the rate at which pressure rises during compression so that
the work required to compress and discharge the contents of the
cylinder is less than that in the isentropic case. Since compressor
efficiency is defined inversely from that of expander efficiency
(i.e., reversible/actual instead of actual/reversible) heat removal
can once again result in theoretical performance greater than
unity.
Although heat transfer can (depending on the direction of heat
flow) skew efficiency measurements upward, no measurements were
made to determine the degree to which heat was added during the
expansion/compression process versus during the exhaust process.
Heat added during the exhaust process does not alter the
theoretical work. For this reason, isentropic efficiency was used
as the measure of device performance despite its potential to
overestimate.
The evaluation apparatus (including expander/compressor and sealed
chamber) was set up on a bench top. The apparatus comprised a
sealed glass enclosure that contained the expander/compressor.
Flow-loop components were installed outside the glass enclosure and
connected through the glass enclosure to the expander/compressor.
The enclosure also contained a miniature dynamometer coupled to the
expander/compressor. The apparatus allowed testing of the
expander/compressor with either a gaseous working fluid, such as
nitrogen gas, or a two-phase refrigerant (such as HFE-7000). A
16-channel computer data-acquisition system was used for collecting
the process data used for calculating or otherwise determining the
performance of the expander/compressor. EES was used for
calculating all performance results obtained using HFE-7000 as the
working fluid.
Certain components of the evaluation apparatus 380 are shown in
FIG. 23, showing the glass enclosure 382 housing the
expander/compressor 384 (comprising expander 384a and compressor
384b). Also contained in the enclosure 382 was a motor 386 for
driving the expander/compressor 384 and a torque sensor 388 coupled
to the motor 386. Outside the enclosure 382 were a pump 390, a
coriolis-based mass-flow controller 392, a vaporizer 394, a needle
valve 396, an evaporator 398, and a thermal-based mass-flow
controller 400. The pump 390 was a magnetically coupled gear pump
(Micropump, Vancouver, Wash.). The coriolis-based mass-flow
controller 392 (Brooks Instruments "Quantim" type QMAC-003K,
Hatfield, Pa.) was used for measuring inlet flow to the expander
384a. The vaporizer 394 comprised an electrically heated copper
block with a series of internal fluid passageways. To promote rapid
heat transfer, a first passageway contained an aluminum plug about
1 mm smaller in diameter than the diameter of the passageway, which
forced liquid to flow along heated walls of the passageway. By
wrapping a fine-mesh screen around the plug, problems with slug
flow were eliminated. Inlet pressure to the expander 384a was
measured using a high-temperature, absolute-pressure sensor 404
(Endevco model 8540-200, San Juan Capistrano, Calif.). For the
compressor 384b, a simple tube-in-fin heat-exchanger was used as
the evaporator 398, and inlet flow was measured using the
thermal-based mass-flow controller 400 (MKS model 0558A-100L-SB,
Wilmington, Mass.). Inlet pressure to the compressor 384b was
measured using an absolute pressure sensor 406 (Omega PX302-030AV,
Stamford, Conn.). Respective temperatures in the expander 384a and
compressor 384b were measured using ungrounded Type-K thermocouples
408 (Omega KQMSS-062U-6, Stamford, Conn.). The motor 386 was a
90-Watt motor (Maxon model 948931, Sachseln, Switzerland) mounted
in a cradle 402 connected to the torque sensor 388. The torque
sensor 388 was a 30 oz-in static-reaction torque sensor, Omega type
TQ202-30Z, Stamford, Conn. The resulting assembly functioned as a
dynamometer. As the motor 386 applied torque to the shaft 384c of
the expander/compressor 384, the reaction torque on the cradle 402
was measured by the torque sensor 388.
For testing the expander 384a using compressed nitrogen as a
working fluid, the glass enclosure 382 was left open and the pump
390 and coriolis-based mass-flow controller 392 were replaced with
a nitrogen tank, pressure regulator, and thermal-based mass-flow
controller (MKS model 179A24CS3BM, Wilmington, Mass.), not shown.
For testing the compressor 384b with air as a working fluid, the
needle valve 396 and evaporator 398 were replaced with a
pressure-reducing regulator (not shown) that had an inlet open to
the atmosphere.
Pressure-volume (P-V) curves from the expander 384a and compressor
384b were obtained using respective piezo-resistive pressure
transducers (Endevco types 8530C-50 and 8530C-100, respectively,
San Juan Capistrano, Calif.) configured for measuring in-cylinder
pressure. A digital encoder (US Digital type E2, Vancouver, Wash.)
was attached to the shaft 384c of the expander/compressor 384 for
measuring rotational position of the shaft. Data from these
components were recorded by the computer programmed with the
LabVIEW program that used the kinematic relationship between the
shaft position and piston positions to convert angle data to
cylinder-volume data. In addition to displaying P-V curves in real
time, the program integrated the P-V curves to determine respective
work done on the piston face during each revolution of the shaft.
By comparing this P-V work to the work measured at the shaft 384c,
frictional losses could be determined. The ratio of these work
values was termed the "transmission efficiency."
FIG. 24 shows the isentropic efficiency of the expander 384a
operating on compressed nitrogen at room temperature and
discharging the nitrogen to atmospheric pressure. Efficiency was
strongly affected by pressure and only weakly affected by operating
speed of the expander 384a. Maximum efficiency was 78% at 1500 rpm
and 100 psia inlet pressure (a pressure-ratio of 6.8). FIG. 25
shows the transmission efficiency (shaft work relative to work done
on the piston face) of the expander 384a. These graphs reveal that
power loss is caused largely by friction. At an inlet pressure of
nearly 80 psia, transmission efficiency equals the isentropic
efficiency, indicating that friction accounts for all expander
losses. This is explained by the earlier-mentioned boost in
performance caused by heat-transfer (i.e., small thermo-fluid
losses are balanced by heat input to the expander 384a from the
environment).
FIG. 26 shows P-V curves recorded from one expander cylinder
operating at 1500 rpm shaft velocity. The graphs reveal a slight
over-expansion when the intake pressure was less than 100 psia,
which was attributed to a slight stickiness of the seat material
(silicone rubber) used in the exhaust valves of the expander, which
prevented the exhaust valves from popping open until after the
cylinder pressure was about 5 psia below exhaust pressure.
FIG. 27 shows the isentropic efficiency of the expander 384a
operating with HFE-7000 as the working fluid. Inlet pressures
varied from 50 to 80 psia, and operating speeds ranged from 500 to
2500 rpm. Although inlet temperature was targeted at 125.degree.
C., some heat loss from the conduits between the vaporizer 394 and
the expander 384a resulted in a slightly lower inlet temperature.
This heat loss can be largely prevented using insulation. Exhaust
pressure varied, according to the saturation pressure of the liquid
in the enclosure 382, from 10.4 to 11.8 psia. The maximum
efficiency was 66% at 1000 rpm and 55 psia inlet pressure. Since
HFE-7000 (molecular weight=200 g/mol) has greater density than
nitrogen, mild pressure drops occurred across the inlet and exhaust
ports of the expander/compressor 384, which yielded lower
efficiency at higher speeds using HFE-7000. At the lowest speed,
some heat loss occurred in the inlet conduit. For the same intake
pressure, up to twice the mass flow per revolution was observed at
500 rpm compared to higher speeds.
FIG. 28 shows P-V curves obtained from one cylinder of the expander
384a at constant inlet pressure and varying speed. The graph shows
how the area inside the curve (which is proportional to the work
done on the piston) decreases as pressure drop increases with
speed. Condensation also affected the data obtained at 500 rpm. As
the pressure drops during expansion, the small amount of liquid
that is present evaporates and raises the pressure above the
higher-speed curves. Thus, more work is produced per revolution at
lower speeds than at higher speeds.
FIG. 29 shows isentropic efficiency of the compressor 384b
operating with room-temperature air, as a working fluid, drawn
through a pressure-reducing regulator. The compressor 384b
discharged the air to the ambient atmosphere at atmospheric
pressure. Maximum efficiency was 69%, which occurred at 2000 rpm
and 6 psig of vacuum (a pressure ratio of 1.68). FIG. 30 shows the
efficiency of transmission calculated from the same data. Unlike
the expander 384a, in which friction causes most of the losses in
that component, the largest losses in the compressor 384b were
fluid losses such as piston leakage. At an intake pressure of 8 psi
of vacuum, the mass-flow per revolution was 2.5 mg/rev at 2500 rpm
versus 1.7 mg/rev at 500 rpm. This was consistent with gas leakage
into cylinder during the intake stroke. The slower rotational
period at 500 rpm allows more gas to leak in so that less is drawn
in from the intake port.
FIG. 31 shows P-V curves recorded from one cylinder of the
compressor 384b. In the figure, leakage effects are evident on the
compression portion of the curves where the pressure rises more
rapidly at 500 rpm than at higher speeds.
The isentropic efficiency of the compressor 384a operating on
HFE-7000 vapor is shown in FIG. 32, in which the operating speed of
the compressor was varied from 500 to 2500 rpm. The exhaust
pressure varied, according to the saturation pressure of the liquid
in the enclosure, from 10.8 to 11.1 psia. The intake pressure was
varied to obtain 2, 4, or 6 psi of vacuum relative to the exhaust
pressure. Because the mass-flow at 2 psi vacuum and 2500 rpm
exceeded the capacity of the flow meter, this point is omitted. The
maximum efficiency was 70% at 1500 rpm and 4 psi of vacuum. At
lower speeds and greater inlet vacuum, leakage around the piston
seals caused reduced efficiency, as shown by the 500-rpm data. As
speed increased, the amount of leakage was reduced in proportion to
the revolution period.
At lesser inlet vacuum and higher speed, efficiency was reduced by
pressure drops in the intake and exhaust ports. FIG. 33 shows P-V
curves recorded from one cylinder with 6 psi inlet vacuum. In the
figure, the pressure drop through the exhaust is evident by the
pressure above 10 psia at the top of the curve, while the pressure
drop through the inlet is evident by the pressure below 4 psia at
the bottom of the curve. These pressure drops result in added work
done by the piston for the same mass of gas compressed. As a
result, the efficiency drops as pressure-drop increases with speed.
The impact on efficiency is greatest for a small intake vacuum
where the added work to overcome the pressure drops is a greater
fraction of the overall work.
Another series of tests were performed on the integrated
expander/compressor using the test apparatus 450 diagrammed in FIG.
34, configured to perform an energy balance on an operating
integrated expander/compressor 452. The integrated
expander/compressor 452 was placed in the test chamber 454 located
in the upper right corner of the diagram. A small-scale dynamometer
was used to monitor the shaft power generated by the expander 452a
and provided a measure of the power input needed by the compressor
452b. The dynamometer comprised a cradled motor 456 and a
high-accuracy static torque sensor 458. Other notable components of
the test apparatus 450 included a cold-side evaporator 460, a
hot-side vaporizer 462, and various temperature-, pressure-, and
flow-measurement devices required for monitoring of system
performance. The test chamber 454 doubled as a condenser. Flow-rate
measurements were key for determining overall cycle performance and
for calculating component efficiencies for the expander 452a and
compressor 452b. A mass-flow meter/controller 464 was used for
determining mass-flow rates of vapor, and a high-accuracy
coriolis-type flow meter 466 was used for measuring liquid flow at
the inlet to the vaporizer 462. All measurements were collected by
a computerized data-acquisition system that allowed performance
data to be displayed on a computer screen in real-time. The
data-acquisition system also displayed pressure-volume (P-V) curves
for the expansion and compression processes occurring within the
cylinders of the expander/compressor 452, which was very important
for assessing sealing and valve operations.
The expander 452a was configured as described in the first
representative embodiment, with cylinders made from hardened
stainless steel for wear resistance and low friction. The cylinder
head was aluminum alloy, and the valve block was made of PEEK. A
silicone gasket formed a seal between the cylinder head and valve
block. The pistons were capped with PPS to form the piston running
surface. The pistons were each 0.5-inch in diameter with a stroke
of 0.48 inch, for a total displacement of 0.189 in.sup.3. Fixed
valve timing gave the expander a volume-expansion ratio of
approximately 3.1. The compressor 452b was configured as described
in the first representative embodiment, with the two pistons each
having a diameter of one inch. The compressor cylinders were
provided with polymeric running surfaces made from PPS, and the
sealing rings (including lip seals as described below) of the
pistons were made of a Delrin-based composite with Teflon
added.
The inlet and exhaust valves of the expander 452a were
spring-loaded poppet-type valves actuated by spring-loaded tappets
mounted within the pistons, as discussed above. The inlet and
exhaust valves of the compressor 452b were passive flapper valves
operating on the inherent pressure differential across each valve,
as discussed above. Improved "breathing" of the compressor
cylinders was accomplished by increasing the porting of the valves.
Also, 0.003-inch thick spring steel was used for the flappers.
The respective cylinders in the expander and compressor utilized
lip seals, as discussed above, to minimize cylinder leakage. By
using a high-temperature polymeric material (PPS) to form the lip
and piston running surfaces, the expander could be operated without
lubrication. Typically, the outside diameter of an expander piston
was approximately 100 micrometers less than the inside diameter of
its cylinder. The piston seals in the compressor 452b were similar,
although no high-temperature polymeric material was needed, so a
Teflon-filled Delrin (having very good sealing capability at the
lower temperatures encountered in the compressor) was used. (A
factor in the consideration of using softer material for cylinder
sealing is the pressure differential used for achieving sealing
action. This pressure differential is orders of magnitude lower in
the compressor than in the expander, which allows the use of a
softer material in the compressor for achieving good contact of the
lip against the cylinder wall during operation.) PPS is a good
alternative material for making sliding surfaces of the compressor
pistons. Although use of PPS is not strictly required for the
compressor pistons, it has good frictional characteristics and also
is compatible with many working fluids (e.g., no swelling of PPS
was evident upon exposure to HFE-7000).
Testing described above showed that the particular embodiment of
the expander that was tested exhibited an isentropic efficiency of
up to 80% using room-temperature nitrogen as the working fluid. The
following testing was directed to an expander/compressor configured
to operate with a refrigerant working fluid at elevated
temperatures (e.g., up to 125.degree. C.). Results are set forth
for the compressor operating on nitrogen and refrigerant and for
the expander operating with heated nitrogen.
FIG. 35 shows P-V curves obtained with the compressor operating at
500 rpm, with air being used as the working fluid. Each curve is a
loop and proceeds counter-clockwise. The substantially horizontal
portions extending across the bottoms of the respective curves
correspond to inlet. The small undulations are due to flow-induced
oscillation of the flapper valves. At the bottom right, the inlet
valve closes as the piston reaches BDC. As the piston begins to
travel upward, the cylinder charge is compressed until the cylinder
pressure reaches the exhaust pressure (atmospheric in the test
setup). The exhaust valve then opens and the cylinder charge is
expelled as the piston moves to TDC. As the piston moves back down,
the exhaust valve closes and the small amount of gas contained in
the clearance space is expanded until the pressure drops to the
intake pressure.
While these curves are nearly ideal, overall performance was
reduced slightly by frictional losses between the piston and
cylinder. FIG. 36 shows the efficiency of the compressor for a
range of operating speeds and exhaust-to-inlet pressure ratios. At
low speeds the compressor exhibited 65% efficiency. At higher
speeds the efficiency was reduced by pressure drops through the
inlet and exhaust valves. This flow loss can be reduced by using
valves having larger ports.
FIG. 37 shows the effect, upon the P-V behavior of the compressor,
of using a fluorocarbon refrigerant as the working fluid. When the
incoming vapor has a temperature near that of the cylinder walls,
the vapor tends to condense on the walls during the compression
stage. This condensed liquid re-evaporates during the subsequent
expansion stage and results in the s-shaped "condensing" curve. As
a result the compressor actually pumps very little refrigerant. If
the incoming vapor is heated to a temperature above the temperature
of the cylinder walls, condensation is avoided and more vapor is
moved through the cylinders.
The higher density (seven times greater than air) of the vapor of
the fluorocarbon refrigerant caused a greater pressure drop through
the valves than experienced with air or nitrogen as a working
fluid. The greater pressure drop resulted in a non-condensing curve
exceeding 14 psia during the exhaust stroke, more than 2 psi higher
than the exhaust pressure of 11.7 psia. Both phenomena, namely the
pressure drop across the valves and condensation, would be reduced
by using larger ports in the valves and using more insulation,
respectively.
FIG. 38 shows P-V curves obtained with the expander 452a operating
at 1500 rpm while inlet pressure was varied from 35 psia to 85
psia. Each curve starts at the left and runs clockwise in a loop.
The inlet phase is first and is indicated by the relatively
straight linear sections extending across the tops of the loops. At
about 0.025 in.sup.3, the inlet valve closes and expansion begins.
The forced opening of the exhaust valve at the end of the expansion
is indicated by the sudden drop in pressure near 0.09 in.sup.3. The
exhaust stroke corresponds to the linear portions extending across
the bottoms and ends of the curves when the exhaust valve closes at
0.015 in.sup.3. A slight recompression occurs before the inlet
valve opens just before TDC.
FIG. 39 shows the calculated isentropic efficiency of the expander
as a function of speed for inlet pressures ranging from 35 psia to
85 psia (20 psig to 70 psig). At the highest pressure, the
isentropic efficiency reaches 70%. At lower pressures,
overexpansion of the gas can result in negative cylinder pressures
before the exhaust valve opens.
Any of various modifications to the system 10 are possible,
depending at least to some extent on the overall size of the system
and the cooling application to which the system will be
applied.
As a first example, a thermoelectric converter ("TEC") can be
disposed between the combustor 20 and the boiler 22 to supply
electrical power for certain tasks and at certain times. For
starting the system 10, certain fluid paths (e.g., conduits
connected to the expander 16 as well as the expander itself)
desirably are pre-heated to prevent condensation. Such pre-heating
can be achieved using one or more heaters driven from a battery
that is recharged with power generated by the TEC. The film heaters
desirably are flexible, high-temperature ribbon heaters that are
capable of withstanding temperatures up to 200.degree. C. A TEC
also can be useful as an auxiliary source of electrical power for
use in driving fans (e.g., for the combustor 20 and/or the
condenser 26) and/or a small pump for fuel delivery from the tank
34 to the combustor 20. Currently available TECs have relatively
low efficiency, which would allow most of the heat from the
combustor 20 to pass through the TEC to the boiler 22. The TEC
would operate at about 5% efficiency with a hot side at
approximately 400.degree. C. to 800.degree. C., while the cool side
of the TEC would operate at or near the temperature of the boiler
22 (e.g., approximately 200.degree. C.). Thus, electrical power for
"balance of plant" and startup tasks, for example, can be generated
readily.
With respect to a second example, for miniaturized systems 10, the
expander 16 and compressor 18 desirably maintain high efficiency
while operating at a sufficiently high mass-flow rate to produce a
desired rate of cooling, taking into account disparate volumetric
flow rates that could exist between the flow of working fluid in
the power cycle 12 and refrigeration cycle 14. One way in which to
address these criteria is to configure the expander as a
radial-flow turbine that is coupled directly to a radial-flow
compressor. This configuration can be more practical (than the
piston configuration described above) for larger systems, whereas
the piston configuration was observed to be more practical in most
instances for miniaturized systems. For example, with smaller
systems employing rotary expanders and compressors, small-diameter
turbine rotors (e.g., diameter of one inch) typically exhibit very
high rotational velocities, and sealing between the rotor and its
housing is difficult to achieve with small rotors. Hence,
positive-displacement configurations for the expander 16 and
compressor 18 are desirable for small-scale systems (e.g.,
performing less than 1 kW of cooling). Positive-displacement
configurations (e.g., pistons operating in cylinders) also tend to
pose less problems with valving of fluids into and out of the
components.
* * * * *
References